KR102234665B1 - Organic rankin cycle and liquefied gas re-gasification and power generation system the same - Google Patents

Abstract

A liquefied gas regasification and cold heat power generation system is disclosed. A liquefied gas regasification and cold heat power generation system according to an embodiment of the present invention includes a liquefied gas transfer line for vaporizing the liquefied gas and sending it to a customer; A heating medium circulation line through which a heating medium for vaporizing the liquefied gas is circulated; An evaporator installed in the heat medium circulation line and vaporizing the heat medium by heat exchange with the heat medium; A vaporizer installed in the heating medium circulation line and regasifying the liquefied gas of the liquefied gas transfer line by using thermal energy and latent heat of the vaporized heating medium; An expansion turbine installed between the evaporator and the vaporizer in the heating medium circulation line and generating power by the power of the heating medium vaporized by the evaporator; A first heat exchanger for heating the liquefied gas regasified by the vaporizer to a temperature required by the customer; The first heat exchanger may receive a part of the heat medium vaporized by the evaporator through a first heat source supply line.

Description

Liquefied gas regasification and cold heat generation system {ORGANIC RANKIN CYCLE AND LIQUEFIED GAS RE-GASIFICATION AND POWER GENERATION SYSTEM THE SAME}

The present invention relates to a liquefied gas regasification and cold heat power generation system, and more particularly, to a liquefied gas regasification and cold heat power generation system that generates power by using the cold heat of a heat medium heat-exchanged with the liquefied gas during the regasification process of the liquefied gas. .

In recent years, as environmental regulations are strengthened, demand for fuels such as natural gas that emit less environmental pollutants is increasing. In order to supply natural gas to consumers, a system for regasifying liquefied natural gas stored in a liquefied state in a liquefied gas storage tank is required.

The conventional liquefied gas regasification system regasifies liquefied natural gas into high-temperature and high-pressure natural gas by exchanging heat source such as seawater and cryogenic liquefied natural gas (LNG). The large-scale cold heat absorbed through the high-temperature heat source cannot be recovered and is discarded into the sea area or the source of the heat source.

The present invention provides a liquefied gas regasification and cold heat power generation system that efficiently powers electricity by using cold heat absorbed as a heat source when the liquefied gas is regasified.

The present invention provides a liquefied gas regasification and cold heat power generation system capable of more flexibly controlling a liquefied gas temperature condition.

The problem to be solved by the present invention is not limited to the problems mentioned above. Other technical problems not mentioned will be clearly understood by those of ordinary skill in the art from the following description.

According to an aspect of the present invention, a liquefied gas transfer line for vaporizing the liquefied gas and sending it to a customer; A heating medium circulation line through which a heating medium for vaporizing the liquefied gas is circulated; An evaporator installed in the heat medium circulation line and vaporizing the heat medium by heat exchange with a heat source; A vaporizer installed in the heating medium circulation line and regasifying the liquefied gas of the liquefied gas transfer line by using thermal energy and latent heat of the vaporized heating medium; An expansion turbine installed between the evaporator and the vaporizer in the heating medium circulation line and generating power by the power of the heating medium vaporized by the evaporator; A first heat exchanger for heating the liquefied gas regasified by the vaporizer to a temperature required by the customer; The first heat exchanger may be provided with a liquefied gas regasification and cold heat power generation system in which a part of the heat medium vaporized by the evaporator is supplied through a first heat source supply line.

In addition, the first heat source supply line may be branched at a front end of the expansion turbine to pass through the first heat exchanger and then connected to the vaporizer.

In addition, it may further include a generator that generates power using the pressure of the vaporized liquefied gas discharged from the first heat exchanger.

In addition, it may further include a second heat exchanger for heating the vaporized liquefied gas discharged from the generator to a temperature required by the customer.

In addition, the second heat exchanger receives a portion of the heat medium vaporized by the evaporator through a second heat source supply line; The second heat source supply line may be branched from a front end of the expansion turbine to pass through the second heat exchanger and then connected to the vaporizer.

In addition, it may further include a second bypass line for supplying the gasified liquefied gas discharged from the first heat exchanger to the customer by bypassing the generator and the second heat exchanger.

In addition, a first bypass line is connected between the evaporator and the vaporizer in parallel with the expansion turbine, the heat medium bypassing the expansion turbine and flowing into the vaporizer; And a first bypass valve installed on the first bypass line.

In addition, when the expansion turbine is damaged, the function of the expansion turbine is restricted, or the flow rate of the heat medium exceeds the allowable capacity of the expansion turbine, a controller may be further included to open the first bypass valve.

In addition, the controller may control the opening degree of the first bypass valve according to the measured values of the temperature and pressure of the heat medium discharged from the evaporator.

According to an embodiment of the present invention, it is possible to efficiently generate electricity by using cold heat absorbed as a heat source during regasification of liquefied gas.

According to an embodiment of the present invention, it is possible to more flexibly control and supply the liquefied gas temperature condition to the temperature required by the customer.

The effects of the present invention are not limited to the above-described effects. Effects not mentioned will be clearly understood by those of ordinary skill in the art from the present specification and the accompanying drawings.

1 is a block diagram of a liquefied gas regasification and cold heat power generation system according to an embodiment of the present invention.
2 is a graph showing a temperature change according to a heat flow amount of a mixed refrigerant used in a liquefied gas regasification system of a ship according to an embodiment of the present invention.
3 is a diagram showing the safety class classification criteria for refrigerants.
4 is a graph showing temperature change characteristics according to the heat flow amount of the mixed refrigerant for each mixing ratio of the mixed refrigerant used in the liquefied gas regasification system of a ship according to an embodiment of the present invention.
5 is a graph illustrating a refrigerant characteristic curve showing the state of the heat medium according to the enthalpy and pressure of the heat medium.
6 is a block diagram of a liquefied gas regasification and cold heat power generation system according to another embodiment of the present invention.
7 is a block diagram of a liquefied gas regasification and cold heat power generation system according to another embodiment of the present invention.
8 is a block diagram of a liquefied gas regasification and cold heat power generation system according to another embodiment of the present invention.

Other advantages and features of the present invention, and a method of achieving them will become apparent with reference to embodiments to be described later in detail together with the accompanying drawings. However, the present invention is not limited to the embodiments disclosed below, and the present invention is only defined by the scope of the claims. Even if not defined, all terms (including technical or scientific terms) used herein have the same meaning as commonly accepted by universal technology in the prior art to which this invention belongs. A general description of known configurations may be omitted so as not to obscure the subject matter of the present invention. In the drawings of the present invention, the same reference numerals are used as much as possible for the same or corresponding configurations. In order to help the understanding of the present invention, some configurations in the drawings may be somewhat exaggerated or reduced.

The terms used in the present application are only used to describe specific embodiments, and are not intended to limit the present invention. Singular expressions include plural expressions unless the context clearly indicates otherwise. In the present application, terms such as "comprise", "have" or "have" are intended to designate the existence of features, numbers, steps, actions, components, parts, or combinations thereof described in the specification. It is to be understood that the possibility of the presence or addition of other features, numbers, steps, actions, components, parts, or combinations thereof, or any further features, is not preliminarily excluded.

1 is a block diagram of a liquefied gas regasification and cold heat power generation system according to an embodiment of the present invention.

The liquefied gas regasification and cold heat power generation system of a ship according to an embodiment of the present invention uses a mixed refrigerant in which refrigerants having a boiling point difference of at least 20°C or more are mixed, and between the dew point temperature and the boiling point temperature in a pressure range of 10 to 20 barg. By vaporizing the liquefied gas using a mixed refrigerant having a difference of 10° C. or more, the temperature is increased during a phase change process in which the mixed refrigerant is vaporized in the evaporator, and the vaporization efficiency of the liquefied gas can be increased.

Referring to FIG. 1, the liquefied gas regasification and cold heat power generation system 100 according to the present embodiment recovers liquefied gases such as liquefied natural gas (LNG) and liquefied petroleum gas (LPG). In order to increase power generation efficiency by supplying fuel gas such as natural gas (NG) and petroleum gas to customers, and generating power by using the cold heat of the liquefied gas in the heat exchange process to regasify the liquefied gas. Can be provided.

In one embodiment, the liquefied gas regasification and cold heat power generation system 100 includes a liquefied gas transfer line 112, a vaporizer 120, a heat medium circulation line 130, a pump 140, a heat medium tank 150, an evaporator ( 160), an expansion turbine 170, and a first heat exchanger 180. Here, it may be referred to as an organic Rankine cycle that is composed of an evaporator, an expansion turbine, a vaporizer, a heat medium tank, and a pump in order on the heat medium circulation line through which the heat medium is circulated.

The liquefied gas transfer line 112 may be provided to receive liquefied gas, vaporize the supplied liquefied gas, and send it to a consumer. The consumer may be, for example, a gas turbine, but is not limited thereto.

For example, a heat medium (mixed refrigerant) in which two or more non-combustible refrigerants are mixed may be circulated in the heat medium circulation line 130. In an embodiment, the ship's liquefied gas regasification system 100 may be configured with a single heat medium circulation line 130.

The heat medium circulates the pump 140, the evaporator 160, the expansion turbine 170, and the carburetor 120 through the heat medium circulation line 130. In order to increase the temperature of the mixed refrigerant in the process of vaporizing the mixed refrigerant in the evaporator 160 and changing the phase from liquid to gas, the heating medium is provided as a refrigerant mixed with refrigerants having a boiling point difference of at least 20°C, and mixed at the same time. The mixing ratio of the refrigerants may be set so that the difference between the dew point temperature and the boiling point temperature of the mixed refrigerant is 10°C or more in the 10 ~ 20 barg pressure (vapor pressure) section of the refrigerant.

The high-pressure booster pump 110 may pressurize and supply cryogenic liquefied gas from a liquefied gas storage tank or a buffer tank to the liquefied gas transfer line 112 in consideration of the pressure required at the place of use. The liquefied gas pressurized by the high-pressure booster pump 110 may be vaporized and heated through the vaporizer 120 to be supplied to a customer.

A heat medium for vaporizing liquefied gas is circulated in the heat medium circulation line 130. The heat medium may be, for example, a fluid and/or a mixture thereof shown in Table 1 below.

ethane Propane I-butane N-butane R13B1 R143a R134a R152a R22 R12 R13 R125 R500 R502 R507 R114 R14 R23 R744

The heat medium circulates through the heat medium tank 150, the pump 140, the evaporator 160, the expansion turbine 170, and the carburetor 120 in sequence along the heat medium circulation line 130.

The vaporizer 120 regasifies the liquefied gas of the liquefied gas transfer line using the thermal energy and latent heat of the heating medium.

The heat medium tank 150 stores the heat medium discharged from the vaporizer 120. The heat medium tank 150 absorbs the pressure change of the heat medium according to the operating conditions, and maintains the heat medium in a set temperature range so that the heat medium can be operated in a predetermined pressure range. The heat medium tank 150 may be provided as an expansion tank. The heat medium tank 150 absorbs the pressure change of the heat medium according to the operating conditions, and maintains the heat medium recovered to the heat medium tank 150 in a set temperature range so that the heat medium tank 150 can be operated in a predetermined pressure range. Two or more types of liquefied refrigerants may be stored in the heat medium tank 150.

The pump 140 pressurizes the heat medium discharged from the heat medium tank 150 and circulates it in the heat medium circulation line 130. The pump 140 may be installed at a rear end of the heat medium tank 150 and a front end of the evaporator 160, and may be configured to pressurize the heat medium and supply it to the evaporator 160.

A first pressure sensor P1 and a first temperature sensor T1 are installed in the heat medium circulation line 130 between the vaporizer 120 and the pump 140. The first pressure sensor P1 and the first temperature sensor T1 measure the pressure and temperature of the liquid heat medium discharged from the vaporizer 120.

The evaporator 160 vaporizes the liquid heat medium discharged from the vaporizer 120 and circulating through heat exchange with a heat source. The heat source may be provided by seawater, steam, waste heat from an engine in a ship, or the like. For example, the evaporator 160 may evaporate the heat medium by heat exchange with a heat source supplied through the seawater line 90. The heat source may be supplied to the evaporator 160 through the seawater line 90 by a seawater pressure pump (not shown). The inlet temperature of the evaporator of the heat source may be about 14° C., and the discharge temperature may be about 7 to 8° C. The evaporator 160 may be manufactured to withstand the pressure of the heating medium (eg, 10 to 20 barg vapor pressure).

When the evaporator 160 is made of, for example, a plate type heat exchanger (PHE) made of titanium, in order to use titanium with a thickness of 0.6 mm or less as a heat exchange plate, the design pressure (vapor pressure) of the heat medium is 16. It should be below barg. If a titanium plate having a thickness of 0.7mm or more is used to increase the design pressure, the cost of equipment is greatly increased because the price increases by more than 10 times compared to a plate having a thickness of 0.6mm.

In the case of the mixed refrigerant, the dew point and the boiling point of the mixed refrigerant are changed according to the type of the mixed refrigerants and the mixing ratio of the refrigerants, thereby affecting the temperature change characteristics of the mixed refrigerant when the phase changes, as well as in the evaporator 160. During the phase change process, the pressure (vapor pressure) of the mixed refrigerant also changes. Accordingly, it is necessary to select the types of refrigerants and set the mixing ratio (weight ratio) so that desirable temperature change characteristics can be realized in the phase change section of the mixed refrigerant while minimizing the equipment cost of the evaporator 160.

In an embodiment of the present invention, in order to increase the temperature of the mixed refrigerant during the phase change of the heat medium in the evaporator 160, the liquefied gas may be regasified using a heat medium in which refrigerants having a boiling point difference of 20° C. or more are mixed. In addition, at a pressure of about 10 to 20 barg, the difference between the dew point temperature and the boiling point temperature of the heating medium is at least 10°C, while the pressure of the heating medium is about 20 barg or less (more preferably Is 16 barg or less), the mixing ratio (weight ratio) of the refrigerants may be set.

For example, the heating medium may include a first refrigerant and a second refrigerant having a boiling point lower than that of the first refrigerant by 20° C. or more. From the viewpoint of increasing the effect of increasing the temperature of the heating medium during the phase change of the heating medium in the evaporator 160, the boiling point of the second refrigerant is preferably 30° C. or more lower than the boiling point of the first refrigerant. 2 The difference in boiling point of the refrigerant is more preferably 50°C or higher.

The expansion turbine 170 is a device that generates electricity from the generator 174 by rotating the turbine 172 when a high-pressure fluid (vaporized heat medium) expands. For example, the expansion turbine 170 is installed between the evaporator 160 and the carburetor 120. The expansion turbine 170 generates power by the power of the heat medium vaporized by the evaporator 160.

A second pressure sensor P2 and a second temperature sensor T2 are installed between the evaporator 160 and the turbine 170. The second pressure sensor P2 and the second temperature sensor T2 measure the pressure and temperature of the heat medium vaporized in the evaporator 160.

A third temperature sensor T3 is installed at the rear end of the vaporizer 120. The third temperature sensor T3 measures the temperature of the liquefied gas vaporized by the vaporizer 120.

The flow rate of the heating medium can be controlled according to the measured values of the first temperature sensor (T1), the first pressure sensor (P1), the second temperature sensor (T2), the second pressure sensor (P2), and the third temperature sensor (T3). I can.

Rankine cycle is applied to the cycle for recovering the cold heat of the liquefied gas. That is, the low temperature liquid heat medium is pressurized through the pump 140, and then vaporized and heated in the evaporator 160 through a heat source. Thereafter, a pressure drop occurs through the expansion turbine 170 for power generation, and thereafter, the low-pressure refrigerant is liquefied again through the vaporizer 120 and circulated.

When the Rankine cycle is applied, since there is a restriction to satisfy the gas temperature and pressure conditions required by the customer, the cold heat that can be recovered from the liquefied gas may be significantly reduced. Since the efficiency of cooling and heat generation depends on the type, pressure and flow rate of the heating medium, it is necessary to maximize the efficiency of cooling and heat generation by optimizing the pressure and flow rate of the pressurized heating medium.

The flow rate and pressure of the heat medium may be determined based on the temperature and pressure of the heat medium discharged from the vaporizer 120, the temperature of the heat source used in the evaporator 160, the pressure point of the heat medium, and the temperature of the gas sent to the customer.

On the other hand, in consideration of the characteristic that the higher the temperature of the heat medium, the higher the pressure that can be operated while maintaining the gaseous state, it is necessary to raise the heat medium to the maximum pressure. However, there is a limitation in that the temperature of the liquefied gas vaporized by heat exchange with the heat medium in the vaporizer 120 must be adjusted to a temperature required by the customer.

Because of this, there may be restrictions on the pressure condition of the heating medium. By providing the first heat exchanger 180 to solve this limitation, the pressure drop of the heating medium is changed to the maximum to increase the amount of heat generation and at the same time You can adjust the gas temperature.

The first heat exchanger 180 receives a part of the heat medium vaporized by the evaporator 160 through the first heat source supply line 182. The first heat source supply line 182 may be branched from the front end of the expansion turbine 170 to pass through the first heat exchanger 180 and then connected to the carburetor 120.

The first heat exchanger 180 receives the heat medium vaporized by the evaporator 160 through the first heat source supply line 182, and heats the vaporized liquefied gas by exchanging heat with the liquefied gas vaporized by the vaporizer 120. It is configured to let.

The heat medium supplied through the first heat source supply line 180 may be designed to be liquefied by heat exchange with the liquefied gas vaporized in the first heat exchanger 180 and supplied to the front end (or the rear end) of the vaporizer 120. .

A fifth pressure sensor (P5) and a fifth temperature sensor (T5) are installed in the first heat source supply line 182, and the flow rate and pressure of the heat medium supplied to the first heat exchanger 220 are determined by a third temperature sensor (T3). ), the fourth temperature sensor T4, the fifth pressure sensor P5, and the fifth temperature sensor T5.

The flow rate ratio of the heat medium passing through the vaporizer 120 and the first heat exchanger 180 may be determined in consideration of the maximum temperature at which the heat medium discharged from the first heat exchanger 180 can change into a liquid phase. The pump 140 or the first flow rate control valve 132 may be controlled according to the pressure and flow rate of the heat medium determined through this process. The pump 140 and the first flow rate control valve 132 may be controlled by the controller 290.

According to this embodiment, by using the heat medium heated by the heat source of the evaporator 160 as a heat source of the first heat exchanger 180, without generating a separate heat source with seawater, steam, atmosphere, and waste heat in the vessel, The system can be miniaturized and the heat source line can be simplified.

2 is a graph showing a temperature change according to a heat flow amount of a mixed refrigerant used in a ship's liquefied gas regasification and cold heat power generation system according to an embodiment of the present invention.

In FIG. 2, the horizontal axis represents the heat flow amount of fluids (seawater, LNG, mixed refrigerant), and the vertical axis represents the temperature of the fluids. The line marked'SW' is a graph of temperature change according to the heat flow of seawater, and the line marked'LNG' is a graph of temperature change according to the heat flow of liquefied natural gas.

According to the present embodiment, refrigerants having a large difference in boiling point are used as the mixed refrigerant, and in the temperature change graph according to the heat flow amount of the mixed refrigerant, the temperature gliding section in which the temperature increases when the phase change vaporized in the liquid state (A) appears, thereby satisfying the condition of the minimum temperature difference (ΔT) compared to the temperature of the liquefied natural gas, and heat exchange between the mixed refrigerant and the liquefied natural gas.

In the case of a single refrigerant, the temperature is constant or the temperature decreases while the phase change occurs, but when refrigerants having a boiling point difference of more than a set value are mixed, the temperature of the mixed refrigerant gradually increases while the phase change of the mixed refrigerant occurs in the evaporator 140. As it rises, a temperature gliding effect occurs at which the mixed refrigerant vaporizes.

When the phase change of the mixed refrigerant, the amount of temperature increase is 2 to 3°C or higher, preferably 10°C or higher, so that refrigerants having a large difference in boiling point are used, and the mixing ratio of the refrigerant having a low boiling point is at least a certain level (mixed refrigerant It is preferable to mix at least 5% by weight). Therefore, it is possible to implement a regasification system that satisfies the condition that the minimum temperature difference between natural gas and the mixed refrigerant is secured, and there is no need to install a trim heater at the rear end of the carburetor, and the thermal energy and latent heat of the mixed refrigerant are sufficiently utilized. Liquefied gas can be vaporized efficiently.

In addition, even when the temperature of the heat source is relatively low, such as seawater, efficient heat transfer to the liquefied gas is possible by using the latent heat of the mixed refrigerant, and the liquefied gas regasification and cold heat power generation system can be simplified and the operation efficiency can be increased. It is possible to increase safety compared to a system using a highly flammable heat transfer medium such as propane by comprising non-flammable refrigerants.

The conventional liquefied gas regasification system is operated by a vaporizer and a trim heater, and in order to operate the refrigerant in a liquid state in the trim heater, the regasifier must be operated at a high pressure to prevent vaporization of the refrigerant. There is a large difference in the operating pressure of the refrigerant between the and trim heaters.

However, according to the present embodiment, the pressure difference in the circulation loop of the heat medium circulation line 130 is small, and since the refrigerant circulates only a single heat exchange loop, it is only necessary to pressurize the pressure consumed in circulation and the head loss to recover the liquefied gas. Energy consumption for fire can be reduced.

As described above, according to the present embodiment, it is possible to maximize the effect of using latent heat of the mixed refrigerant to improve the liquefied gas vaporization performance, to reduce the amount of refrigerant used in the liquefied gas regasification and cold heat power generation system, and also, a trim heater is provided. There is no need to install, and the liquefied gas can be vaporized with a single vaporizer, thus simplifying the system configuration.

In addition, liquefied gas can be regasified by a single heat exchange process without going through two-stage heat exchange, and energy required for the pump and pipe size can be reduced by reducing the circulation flow rate of the mixed refrigerant. The cost can also be reduced.

In addition, according to an embodiment of the present invention, it is possible to implement a liquefied gas regasification system that efficiently utilizes the latent heat of the mixed refrigerant while changing the phase of the mixed refrigerant even in a low seawater design temperature range. In the case of using the evaporative mixed refrigerant, since the latent heat is much larger than the sensible heat, the flow rate of the circulating mixed refrigerant can be reduced, and thus the energy required for circulation of the mixed refrigerant is reduced.

In order to maximize the temperature increase effect during the phase change of the mixed refrigerant as described above, the mixed refrigerant includes a first refrigerant having a boiling point of -40 to -10°C, and a second refrigerant having a boiling point of -60 to -90°C. Can include. A more preferable difference in boiling point between the first refrigerant and the second refrigerant is 50° C. or more.

In order to control the pressure of the mixed refrigerant below the design pressure of the evaporator 160 and at the same time obtain a temperature increase effect when the mixed refrigerant phase changes in the evaporator 160, the second refrigerant has a boiling point lower than that of the first refrigerant. The refrigerant may be mixed with the first refrigerant to have a weight ratio of 0.05 to 0.15 based on the weight of the mixed refrigerant.

The mixed refrigerant may be composed of refrigerants having a non-flammable property, an ozone depletion index of 0, and a global warming index of less than 2000. 3 is a diagram showing the safety class classification criteria for refrigerants. As shown in FIG. 3, the refrigerant is classified into safety grades according to toxicity and flammability.

In an embodiment, the refrigerants of the mixed refrigerant may be selected from refrigerants having a safety class of A1, except for refrigerants having a safety group of B1 to B3 or A2 to A3. That is, a mixed refrigerant can be formed in consideration of only refrigerants having a safety grade having low toxicity and low flammability. In addition, the refrigerants of the mixed refrigerant are preferably selected from eco-friendly refrigerants having an ozone depletion index of 0 and a global warming index of less than 2000 (more preferably less than 1500).

In an embodiment, the mixed refrigerant is tetrafluoroethane (1,1,1,2-tetrafluoroethane, R134a; CH ₂ FCF ₃ ) corresponding to the first refrigerant and carbon dioxide (R744; CO _{2) corresponding to the second refrigerant.} ), including at least two or more refrigerants. Two refrigerants (R134a, R744) are refrigerants classified as safety grades corresponding to A1, are non-toxic and non-flammable, and have an Ozone Depletion Potential (ODP) of 0. In addition, R744 refrigerant has a Global Warming Potential (GWP) of 1, and R134a refrigerant has a GWP of 1430, which is higher than R744 refrigerant, but has a lower value than the GWP standard suggested by various regulations and classification so that it meets the regulations It is suitable for use.

Because R744 is more volatile than R134a, its boiling point is about 50°C or more. R134a has a boiling point of about -26℃, and R744 has a boiling point of about -78℃. Accordingly, when a mixed refrigerant in which two refrigerants R134a and R744 are mixed is used, a temperature gliding effect in which a temperature increases during a phase change due to a large difference in boiling point of the two refrigerants can be sufficiently obtained.

The mixed refrigerant may contain tetrafluoroethane (CH ₂ FCF ₃ , R134a) and carbon dioxide (CO ₂ , R744) in a weight ratio of 85 to 95% and 5 to 15% by weight, respectively. When the R744 refrigerant is contained in an amount of less than 5% by weight, it is difficult to obtain a temperature increase effect when the phase change of the mixed refrigerant due to the high boiling point difference between R744 and R134a.

Conversely, when the mixing ratio of the R744 refrigerant exceeds 15% by weight, the pressure of the mixed refrigerant in the evaporator 160 exceeds 20 barg due to the high content of R744 having high volatility, and withstands a high pressure of 20 barg or more. In order to be able to do so, there may be a problem in that the design thickness of the evaporator 160 must be increased.

4 is a graph showing temperature change characteristics according to the heat flow amount of the mixed refrigerant for each mixing ratio of the mixed refrigerant used in the liquefied gas regasification system of a ship according to an embodiment of the present invention. A mixed refrigerant mixed with CH ₂ FCF ₃ (R134a) and CO ₂ (R744) was used, and the temperature change characteristics of the mixed refrigerant while changing the weight ratio of R744 to 0.05, 0.1, and 0.15, the minimum between the mixed refrigerant and seawater (SW). The proximity temperature, the saturated vapor pressure of the mixed refrigerant in the evaporator, and the flow rate of the mixed refrigerant were measured.

Table 2 shows the minimum proximity temperature (minimum temperature between the mixed refrigerant and seawater) of the mixed refrigerant according to the weight ratio of _{CH 2} FCF ₃ (R134a) and CO _{2 (R744) constituting the mixed refrigerant, and the saturated vapor pressure of the mixed refrigerant (based on 50℃)} And changes in the flow rate of the mixed refrigerant. From Table 1, it can be seen that as the weight ratio of R744 increases, the saturated vapor pressure of the mixed refrigerant in the evaporator increases.

The more volatile of R134a and R744 is R744. The more R744 is mixed, the greater the temperature gliding effect during phase change of the mixed refrigerant, so that the refrigerant circulation flow rate decreases.However, the vapor pressure of the mixed refrigerant in the evaporator 140 increases, making a heat exchanger. May not satisfy the design pressure.

In the offshore platform or ship, there is a limitation that the system must be configured with a limited space compared to the onshore plant. On land, an open-rack vaporizer that vaporizes fluid using air or seawater can be used, but there are restrictions on installing and operating such a large-scale system on a ship. Accordingly, compact equipment such as a plate heat exchanger (PHE) and a printed circuit heat exchanger (PCHE) must be operated, and such equipment has limitations in manufacturing size and price.

In the evaporator 160 used in the circulation cycle of the mixed refrigerant, the performance, price, design pressure, etc. of the heat exchanger are determined according to the thickness of the plate, which is an internal component. In the refrigerant system, the system is operated by setting the system design pressure to the vapor saturation pressure at a specific temperature. Since R744 is more volatile than R134a, the design pressure increases as the mixing composition ratio of R744 increases.

Therefore, in order to satisfy the design pressure of the mixed refrigerant in consideration of the performance and price of the evaporator 160 used in the ship, it is preferable to control the mixing composition ratio of R744 to 15% by weight or less. Therefore, in order to obtain a temperature increase effect during phase change of the mixed refrigerant and to control the pressure of the mixed refrigerant below the design pressure, the mixing ratio of R744 is controlled so that 5 to 15% by weight of R744 is contained in the mixed refrigerant. It is desirable to do it.

When the evaporator is provided as a plate heat exchanger made of a 0.6mm thick tungsten plate to withstand the pressure of the mixed refrigerant (saturated vapor pressure) of 16 barg, the pressure of the mixed refrigerant in the evaporator (saturated vapor pressure) is 16 at an outside temperature of 50°. In order to satisfy the condition of barg or less, it is preferable that the mixing ratio of the second refrigerant R744 having a low boiling point among the mixed refrigerants is 10% by weight or less.

Table 3 shows the change in boiling point temperature and dew point temperature of the mixed refrigerant according to the pressure of the mixed refrigerant when the weight ratio of R134a and R744 is 0.95:0.05 and 0.85:0.15. When the ratio of R744 is high (when the weight ratio of R744 is 15%), the difference between the dew point temperature and the bubble point temperature of the mixed refrigerant increases, and the temperature increase effect increases when the mixed refrigerant phase changes. In addition, it becomes advantageous to use a mixed refrigerant within the limited temperature conditions of seawater and LNG.

When the ratio of R744 is low (when the weight ratio of R744 is 5%), the difference between the dew point temperature and the boiling point temperature of the mixed refrigerant decreases. When the weight ratio of R744 is less than 5%, the difference between the dew point temperature and the boiling point temperature of the mixed refrigerant at a vapor pressure of 10 to 20 barg decreases to less than about 10°C, resulting in a temperature increase effect when the phase change of the mixed refrigerant in the evaporator 140 May decrease. Therefore, it is desirable to keep R744 above 5% by weight.

5 is a graph illustrating a refrigerant characteristic curve showing the state of the heat medium according to the enthalpy and pressure of the heat medium.

1 and 5, the flow rate and pressure of the heat medium flowing into the expansion turbine 170 is such that the heat medium flowing into the expansion turbine 170 does not deviate from the gas-liquid mixing section in which the expansion turbine 170 can be operated normally. In addition, the heat medium discharged from the vaporizer 120 may be controlled to become a liquid. That is, when the heat medium introduced into the expansion turbine 170 is discharged, the pressure applied to the expansion turbine 170 may be limited so that it does not go beyond the operating gas-liquid mixing section.

6 is a block diagram of a liquefied gas regasification and cold heat power generation system according to another embodiment of the present invention.

In describing the embodiment of FIG. 6, duplicate descriptions of components that are the same as or corresponding to the above-described embodiment may be omitted.

The liquefied gas regasification and cold heat power generation system 100a according to the embodiment of FIG. 6 is different from the above-described embodiment in that it further includes a generator 260 and a second heat exchanger 240.

The generator 260 may be an expansion turbine that generates power by directly using the pressure of the vaporized liquefied gas discharged from the heater. The second heat exchanger 240 may heat the vaporized liquefied gas discharged from the generator 260 to a temperature required by a customer.

The second heat exchanger 240 receives a part of the heat medium vaporized by the evaporator 160 through the second heat source supply line 242, and the second heat source supply line 242 branches from the front end of the expansion turbine 170 It is connected to the carburetor 120 after passing through the second heat exchanger (240).

According to the embodiment of FIG. 6, additional power generation is possible using the gas pressure of the first heat exchanger 180 at the rear end according to the pressure required by the customer. At this time, since the temperature drop also occurs according to the pressure drop of the gas in the generator 260, the second heat exchanger 240 heats NG using seawater, steam, atmosphere, and waste heat in the ship and supplies it to the customer. The turbine of the generator 260 may be connected coaxially with the expansion turbine 170 used for cold and heat power generation of liquefied gas.

7 is a block diagram of a liquefied gas regasification and cold heat power generation system according to another embodiment of the present invention.

In describing the embodiment of FIG. 7, redundant descriptions of components that are the same as or corresponding to the above-described embodiment may be omitted.

The liquefied gas regasification and cooling and heat generation system 100b according to the embodiment of FIG. 7 further includes a first bypass line 102, a first bypass valve 104, and a flow control valve 106. There is a difference from the described embodiment.

The first bypass line 102 may be connected in parallel with the expansion turbine 170 between the evaporator 160 and the vaporizer 120. The first bypass line 102 may allow the heat medium to flow into the carburetor 120 by bypassing the expansion turbine 170.

The first bypass valve 104 is installed in the first bypass line 102 and may adjust the flow rate of the heat medium bypassing the expansion turbine 170.

The flow control valve 106 is installed in front of the expansion turbine 170 and controls the flow rate of the heat medium flowing into the expansion turbine 170.

The controller 290 is a first bypass so that the heat medium flowing into the expansion turbine 170 does not deviate from the gas-liquid mixing section in which the expansion turbine 170 can operate normally, and the heat medium discharged from the carburetor 120 becomes liquid. The flow rate and pressure of the heat medium flowing into the expansion turbine 170 may be controlled by controlling the valve 104 and the flow control valve 106.

The controller 290 includes a first temperature sensor (T1), a first pressure sensor (P1), a second temperature sensor (T2), a second pressure sensor (P2), a third temperature sensor (T3), and a seventh temperature sensor ( Based on measured values such as T7), the pump 140, the first flow rate control valve 132, the first bypass valve 104, and the flow control valve 106 may be controlled.

In one embodiment, the controller 290 is the first battery in case the expansion turbine 170 is damaged, the function of the expansion turbine 170 is restricted, or the flow rate of the heat medium exceeds the allowable capacity of the expansion turbine 170. The pass valve 104 can be opened.

The controller 290 determines the vaporization performance of the vaporizer 120 and the power generation efficiency of the expansion turbine 170 in consideration of, for example, the heat medium discharge temperature of the vaporizer 120 based on the refrigerant characteristic curve as shown in FIG. 2. It can be set by calculating the pressurized pressure value of the heating medium for optimization. At this time, the controller 290 limits the optimum pressure point of the heat medium so that the heat medium pressure-enhanced in the expansion turbine 170 satisfies the allowable phase change range of the expansion turbine 170.

In order to optimize the refrigerant flow rate, the amount of NG gas required by the consumer should be considered, and the recovery rate of cold heat in the carburetor 120 must be increased.To this end, the amount of heat supplied from the carburetor 120, which is a section for vaporizing and heating LNG, must be maximized. . However, since the heat medium passing through the vaporizer 120 is drawn into the pump 140 after being discharged from the vaporizer 120, it must be in a liquid state. Accordingly, the controller 290 senses the temperature of the heat medium at the rear end of the vaporizer 120 and adjusts the flow rate of the heat medium so that the temperature becomes liquid under the pressure condition of the vaporizer 120.

In the controller 290, the pressure and flow rate at the rear end of the heat exchanger of the heat medium/heat source, which are optimally calculated through the calculation process by the combination of the characteristics of the fluid and the temperature and pressure at each point, is a frequency controller (VFD; Variable Frequency Drive) connected to the pump 140. ) Or through the first flow control valve 132 installed at the rear end of the pump 140.

When the flow rate of the heating medium circulated for regasification of the liquefied gas exceeds the capacity of the expansion turbine 170, the flow rate of the heating medium is limited due to the allowable capacity of the expansion turbine 170 to provide a sufficient heat source for regasification of the liquefied gas. It will not be able to supply. In order to prevent this, when the expansion turbine 170 is installed in a large capacity, power generation efficiency is degraded because the main use point always exceeds the appropriate circulation amount.

On the other hand, according to the embodiment of FIG. 7, when the expansion turbine 170 is damaged, when the expansion turbine 170 is capable of only some limited functions, or when it is necessary to supply a heat medium exceeding the capacity of the expansion turbine 170 As described above, when there is a need to temporarily increase the flow rate of the heat medium to be supplied to the vaporizer 120, the heat medium is additionally supplied through the first bypass line 102 to facilitate the regasification operation. .

In the case of the heat medium flowing into the carburetor 120 through the first bypass line 102, the energy state is higher than the outlet of the expansion turbine 170 that has recovered kinetic energy, so only a relatively small flow rate is the first bypass line ( Regasification can also be achieved by supplying to the vaporizer 120 through 102).

When the heat medium is bypassed through the first bypass line 102, there may be a difference from the amount of energy transfer used in cold and heat generation. In this case, the controller 290 senses the temperature of the heat medium at the rear end of the vaporizer 120, The flow rate of the heat medium passing through the flow control valve 106 and the first bypass valve 104 is adjusted so that the heat medium at the rear end of the vaporizer 120 maintains a liquid state.

Even when the operation of the expansion turbine 170 is impossible, regasification operation is possible through the first bypass line 102.

8 is a block diagram of a liquefied gas regasification and cold heat power generation system according to another embodiment of the present invention.

In describing the embodiment of FIG. 8, duplicate descriptions of components that are the same as or corresponding to the above-described embodiment may be omitted.

The liquefied gas regasification and cold heat power generation system 100c according to the embodiment of FIG. 8 is different from the previously described embodiment in that it further includes a second bypass line 116 and a second bypass valve 118. .

The second bypass line 116 may allow the vaporized liquefied gas discharged from the first heat exchanger to flow into a customer by bypassing the generator and the second heat exchanger.

It is to be understood that the above embodiments have been presented to aid understanding of the present invention, do not limit the scope of the present invention, and various deformable embodiments from this also fall within the scope of the present invention. The technical protection scope of the present invention should be determined by the technical idea of the claims, and the technical protection scope of the present invention is not limited to the literal description of the claims itself, but a scope that has substantially equal technical value. It should be understood that it extends to the invention of.

100: Liquefied gas regasification and cold heat power generation system
110: high pressure booster pump 112: liquefied gas transfer line
120: carburetor 130: heat medium circulation line
140: pump 150: heat medium tank
160: evaporator 170: expansion turbine

Claims (12)

A liquefied gas transfer line for vaporizing the liquefied gas and sending it to a customer;
A heating medium circulation line through which a heating medium for vaporizing the liquefied gas is circulated;
An evaporator installed in the heat medium circulation line and vaporizing the heat medium by heat exchange with a heat source;
A vaporizer installed in the heating medium circulation line and regasifying the liquefied gas of the liquefied gas transfer line by using thermal energy and latent heat of the vaporized heating medium;
An expansion turbine installed between the evaporator and the vaporizer in the heating medium circulation line and generating power by the power of the heating medium vaporized by the evaporator;
A first heat exchanger for heating the liquefied gas regasified by the vaporizer to a temperature required by the customer;
A generator generating power using the pressure of the vaporized liquefied gas discharged from the first heat exchanger;
And a second heat exchanger for heating the vaporized liquefied gas discharged from the generator to a temperature required by the customer;
The first heat exchanger receives a part of the heat medium vaporized by the evaporator through a first heat source supply line,
Liquefied gas regasification and cold heat power generation system comprising a second bypass line for supplying the vaporized liquefied gas discharged from the first heat exchanger to the customer by bypassing the generator and the second heat exchanger. The method of claim 1,
The first heat source supply line is branched from the front end of the expansion turbine, passes through the first heat exchanger, and is connected to the vaporizer. delete delete The method of claim 1,
The second heat exchanger receives a part of the heat medium vaporized by the evaporator through a second heat source supply line;
The second heat source supply line is branched from the front end of the expansion turbine, passes through the second heat exchanger, and is connected to the vaporizer. The method of claim 1,
A first bypass line connected between the evaporator and the vaporizer in parallel with the expansion turbine and allowing the heat medium to bypass the expansion turbine and flow into the vaporizer; And
Liquefied gas regasification and cold heat power generation system further comprising a first bypass valve installed in the first bypass line. The method of claim 6,
Regasification and cooling of liquefied gas further comprising a controller for opening the first bypass valve when the expansion turbine is damaged, the function of the expansion turbine is restricted, or the flow rate of the heat medium exceeds the allowable capacity of the expansion turbine. Power generation system. The method of claim 7,
The controller is a liquefied gas regasification and cold heat power generation system for controlling the opening of the first bypass valve according to the measured values of the temperature and pressure of the heat medium discharged from the evaporator. delete The method of claim 1,
The heating medium is a mixed refrigerant in which two or more refrigerants having different boiling points are mixed. The method of claim 10,
The mixed refrigerant is a liquefied gas regasification and cold heat power generation system including a first refrigerant having a boiling point of -40 to -10°C and a second refrigerant having a boiling point of -60 to -90°C. The method of claim 10,
The refrigerants include a first refrigerant and a second refrigerant having a boiling point lower than that of the first refrigerant by 20° C. or more,
The second refrigerant is a liquefied gas regasification and cold heat power generation system that is mixed with the first refrigerant to have a weight ratio of 0.05 to 0.15 based on the weight of the mixed refrigerant

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