JP2016125746A

JP2016125746A - Refrigerator or air conditioner and control method for the same

Info

Publication number: JP2016125746A
Application number: JP2014266465A
Authority: JP
Inventors: 展海猪野; Hiromi Ino; 宜三関屋; Yoshizo Sekiya; 敏和寒風澤; Toshikazu Sabuzawa; 明登町田; Akito Machida
Original assignee: Mayekawa Manufacturing Co
Current assignee: Mayekawa Manufacturing Co
Priority date: 2014-12-26
Filing date: 2014-12-26
Publication date: 2016-07-11
Anticipated expiration: 2034-12-26
Also published as: WO2016104147A1; JP6456139B2

Abstract

PROBLEM TO BE SOLVED: To improve a thermal efficiency of a refrigerator or an air conditioner constituting an approximation reverse Ericsson cycle.SOLUTION: This invention forms a reverse Ericsson cycle mainly comprising: an isothermal heat radiation stroke by a condenser; an isothermal heat absorption stroke by an evaporator; an isobaric heat radiation stroke at a liquid region performed by heat-exchanging at a regenerative heat exchanger; and an isobaric heat absorption stroke in a super-heated steam region. A partial stroke performed in the super-heated steam region of the isothermal heat radiation stroke is replaced with heat insulation compression strokes of a plurality of stages by the compressors of a plurality of stages and the isobaric heat radiation strokes of a plurality of stages by a gas cooler and the condenser, the refrigerant of wet state at the isobaric heat absorption stroke is fed into the regenerative heat exchanger. There is provided a control device for performing a controlling operation in such a way that the final point of the isobaric heat radiation stroke is positioned from a sensible heat limit super-cooled point becoming a final point of the isobaric heat radiation stroke when the saturated gas refrigerant is fed into the regenerative heat exchanger at the isobaric heat absorption stroke to a wet limit super-cooled point becoming a final point of the isobaric heat radiation stroke when the final point of a heat insulation expansion stroke by the first expansion means is positioned on a saturation refrigerant liquid line.SELECTED DRAWING: Figure 1

Description

本発明は、冷凍サイクルとして逆エリクソンサイクルを用いた冷凍又は空調装置及びその制御方法に関する。 The present invention relates to a refrigeration or air conditioner using a reverse Ericsson cycle as a refrigeration cycle and a control method thereof.

蒸気圧縮式冷凍サイクルを用いる冷凍又は空調装置は、省エネの観点から熱効率の向上が望まれており、従来から多くの提案がなされている。
元来、２つの異なる温度の熱源を用いて蒸気圧縮式冷凍サイクルを構成する冷凍又は空調装置の熱効率は、サイクルが可逆サイクルであるとき理論上可能な最高値に達する。
実用的な冷凍又は空調装置の熱効率をこの理論値に近づける最良の方法の一つは、蒸気圧縮式冷凍サイクルを理論上の逆エリクソンサイクルに近づけることである。 In a refrigeration or air conditioner using a vapor compression refrigeration cycle, improvement in thermal efficiency is desired from the viewpoint of energy saving, and many proposals have been made conventionally.
Originally, the thermal efficiency of a refrigeration or air conditioner comprising a vapor compression refrigeration cycle using two different temperature heat sources reaches the highest theoretically possible when the cycle is a reversible cycle.
One of the best ways to bring the thermal efficiency of a practical refrigeration or air conditioner closer to this theoretical value is to bring the vapor compression refrigeration cycle closer to the theoretical inverse Ericsson cycle.

理論上の逆エリクソンサイクルは、２つの等温過程と２つの等圧過程とからなり、これら２つの等圧過程は熱交換過程である。実用的な冷凍又は空調装置では、冷媒循環路に直列に設けられた圧縮機、凝縮器、膨張弁及び蒸発器等の冷凍サイクル構成機器と共に、蒸発器出口から圧縮機入口に向かうガス冷媒と凝縮器出口から膨張弁入口に向かう液冷媒とを熱交換させる再生熱交換器とを備えることで、理論上の逆エリクソンサイクルに近似した逆エリクソンサイクルが可能になる。 The theoretical inverse Ericsson cycle consists of two isothermal processes and two isobaric processes, which are heat exchange processes. In practical refrigeration or air conditioners, together with refrigeration cycle components such as compressors, condensers, expansion valves and evaporators provided in series in the refrigerant circuit, gas refrigerant and condensation from the evaporator outlet to the compressor inlet are condensed. By providing the regenerative heat exchanger that exchanges heat with the liquid refrigerant from the outlet to the expansion valve inlet, a reverse Ericsson cycle that approximates the theoretical reverse Ericsson cycle becomes possible.

特許文献１には、前述の近似した逆エリクソンサイクルを用いた冷凍又は空調装置が開示されている。この近似逆エリクソンサイクルでは、等温放熱行程の中で過熱蒸気領域で行われる等温圧縮の部分行程を多段の断熱圧縮行程と多段の等圧放熱行程で置き換え、かつ蒸発器出口から再生熱交換器入口に流入するガス冷媒の乾き度を制御することで、熱効率を向上させることができる。 Patent Document 1 discloses a refrigeration or air conditioner using the above-described approximate inverse Ericsson cycle. In this approximate inverse Ericsson cycle, the partial process of isothermal compression performed in the superheated steam region in the isothermal heat release process is replaced with a multistage adiabatic compression process and a multistage isobaric heat release process, and from the evaporator outlet to the regenerative heat exchanger inlet. Thermal efficiency can be improved by controlling the dryness of the gas refrigerant flowing into the.

特許第４７２６２５８号公報Japanese Patent No. 4726258

再生熱交換器でのガス冷媒と液冷媒との熱交換においては、ガス冷媒の比熱が液冷媒と比べて小さいため、液冷媒側が過冷却不足となる傾向があり、熱効率の向上があまり見込めないという問題がある。
特許文献１に開示された冷凍又は空調装置では、蒸発器から再生熱交換器に流入するガス冷媒の乾き度（湿り度）を制御することで、熱効率を向上させようとするものであるが、蒸発器出口の冷媒状態を正確に適正湿り状態に制御することが難しい。最適な制御点は一点のみであるから、蒸発器出口の冷媒状態が湿り度過剰でも湿り度不足でもサイクルの成績係数は低下する。 In heat exchange between the gas refrigerant and liquid refrigerant in the regenerative heat exchanger, the specific heat of the gas refrigerant is smaller than that of the liquid refrigerant, so the liquid refrigerant side tends to be undercooled, and the improvement in thermal efficiency cannot be expected much. There is a problem.
In the refrigeration or air-conditioning device disclosed in Patent Document 1, it is intended to improve the thermal efficiency by controlling the dryness (wetness) of the gas refrigerant flowing from the evaporator to the regenerative heat exchanger, It is difficult to accurately control the refrigerant state at the outlet of the evaporator to a proper wet state. Since the optimum control point is only one point, the coefficient of performance of the cycle is lowered regardless of whether the refrigerant state at the evaporator outlet is excessively wet or insufficient.

本発明の少なくとも一実施形態は、前述の近似逆エリクソンサイクルを用いた冷凍又は空調装置において、熱効率のさらなる向上を可能にすることを目的とする。 An object of at least one embodiment of the present invention is to enable further improvement in thermal efficiency in a refrigeration or air-conditioning apparatus using the above approximate inverse Ericsson cycle.

本発明の少なくとも一実施形態に係る冷凍又は空調装置は、
（１）冷媒循環路に直列に設けられた複数段の圧縮機、凝縮器、第１膨張手段及び蒸発器と、
前記複数段の圧縮機で各圧縮機の出口と次段の圧縮機の入口とを結ぶ前記冷媒循環路に設けられ、各圧縮機の吐出ガス冷媒を冷却するためのガス冷却器と、
前記蒸発器より前記圧縮機に向かう低圧側ガス冷媒と前記凝縮器から前記第１膨張手段に向かう高圧側液冷媒とを熱交換させるための再生熱交換器と、を備え、
主として前記凝縮器による等温放熱行程と前記蒸発器による等温吸熱行程と、前記再生熱交換器における高圧側液冷媒による等圧放熱行程及び低圧側ガス冷媒による等圧吸熱行程から成る逆エリクソンサイクルを形成すると共に、
前記等温放熱行程のうち過熱蒸気領域で行われる等温圧縮の部分行程が、前記複数段の圧縮機による複数段の断熱圧縮行程と、前記ガス冷却器による複数段の等圧放熱行程とに置き換えられることにより近似的等温行程が形成され、
前記等圧吸熱行程開始点で湿り状態の低圧側ガス冷媒を前記再生熱交換器に導入し、前記等圧吸熱行程開始点で飽和状態の低圧側ガス冷媒が前記再生熱交換器に導入されたときの高圧側液冷媒の等圧放熱行程終了点となる顕熱限界過冷却点から前記第１膨張手段による断熱膨張行程の終了点が蒸発圧力における飽和冷媒液線上に位置するときの前記等圧放熱行程の終了点となる湿り限界過冷却点までの過冷却領域内に、前記等圧放熱行程の終了点が位置するように制御するための制御装置をさらに備えている。 The refrigeration or air conditioning apparatus according to at least one embodiment of the present invention is:
(1) a plurality of stages of compressors, condensers, first expansion means and evaporators provided in series in the refrigerant circuit;
A gas cooler for cooling the discharged gas refrigerant of each compressor, provided in the refrigerant circulation path connecting the outlet of each compressor and the inlet of the next stage compressor in the plurality of stages of compressors;
A regenerative heat exchanger for exchanging heat between the low-pressure side gas refrigerant heading from the evaporator toward the compressor and the high-pressure side liquid refrigerant heading from the condenser toward the first expansion means,
An inverse Ericsson cycle mainly comprising an isothermal heat release process by the condenser, an isothermal heat absorption process by the evaporator, an isobaric heat release process by the high pressure side liquid refrigerant and an isobaric heat absorption process by the low pressure side gas refrigerant in the regeneration heat exchanger is formed. As well as
Of the isothermal heat release process, the partial process of isothermal compression performed in the superheated steam region is replaced with a plurality of stages of adiabatic compression processes by the plurality of stages of compressors and a plurality of stages of isobaric heat release processes by the gas cooler. This forms an approximate isothermal process,
The wet low-pressure side gas refrigerant is introduced into the regeneration heat exchanger at the starting point of the isobaric endothermic stroke, and the saturated low-pressure side gas refrigerant is introduced into the regeneration heat exchanger at the starting point of the isobaric endothermic stroke. When the end point of the adiabatic expansion process by the first expansion means is located on the saturated refrigerant liquid line at the evaporation pressure from the sensible heat limit supercooling point which is the end point of the isobaric heat dissipation process of the high-pressure side liquid refrigerant The apparatus further includes a control device for controlling the end point of the isobaric heat dissipation process to be located in the supercooling region up to the wetness limit supercooling point that is the end point of the heat dissipation process.

図１はＴ−Ｓ線図における逆エリクソンサイクルの概略図であり、図２はモリエル線図における逆エリクソンサイクルの概略図である。
図１、図２中、符号ｘは冷媒の飽和液線、ｙは冷媒の飽和蒸気線、ｔは等温線、Ｔｖは単段圧縮機の吐出温度、図１中の破線ｐは図２中の水平線b’gに対応する等圧線、図２中の破線Ｓは図１中の垂直線bb’に対応する等エントロピ線を示している。理論上の逆エリクソンサイクルはａｂｇｃｄa（実線）で示される。理論上の逆エリクソンサイクルは２つの等温行程（ｄ→ａ及びｂ→ｃ）と２つの等圧行程（ａ→ｂ及びｃ→ｄ）からなる。２つの等圧行程abおよびcdは再生熱交換器における熱交換過程である。 FIG. 1 is a schematic diagram of a reverse Ericsson cycle in a TS diagram, and FIG. 2 is a schematic diagram of a reverse Ericsson cycle in a Mollier diagram.
1 and 2, the symbol x is a saturated liquid line of the refrigerant, y is a saturated vapor line of the refrigerant, t is an isotherm, Tv is a discharge temperature of the single stage compressor, and a broken line p in FIG. 1 is in FIG. The isobaric line corresponding to the horizontal line b′g and the broken line S in FIG. 2 indicate the isentropic line corresponding to the vertical line bb ′ in FIG. The theoretical inverse Ericsson cycle is indicated by abgcda (solid line). The theoretical inverse Ericsson cycle consists of two isothermal strokes (d → a and b → c) and two isobaric strokes (a → b and c → d). The two isobaric strokes ab and cd are heat exchange processes in the regenerative heat exchanger.

部分等温行程（ｂ→ｇ）は理論的等温圧縮行程である。実用サイクルでは断熱圧縮行程（ｂ→b’）と等圧放熱行程（b’→ｇ）とに置き換える必要があり、そのため、圧縮仕事Δが余分に必要となる。等圧行程b’gはガス冷却器によって行われる。
そこで、図中実線ｉで示すように、前記等圧放熱行程のうち過熱蒸気領域で行われる部分行程（等温行程ｂ→ｇ）を、複数段の圧縮機による複数段の断熱圧縮行程と、ガス冷却器及び最終段の顕熱部分を冷却する凝縮器による複数段の等圧放熱行程とに置き換えることで、等温行程（ｂ→ｇ）に近似できる。
前記複数段のサイクルを構成することで、圧縮仕事Δを低減でき、ＣＯＰ（成績係数）の向上が可能となる。 The partial isothermal stroke (b → g) is a theoretical isothermal compression stroke. In the practical cycle, it is necessary to replace the adiabatic compression process (b → b ′) and the isobaric heat radiation process (b ′ → g), and therefore, an additional compression work Δ is required. The isobaric stroke b′g is performed by a gas cooler.
Therefore, as shown by a solid line i in the figure, a partial stroke (isothermal stroke b → g) performed in the superheated steam region in the isobaric heat radiation stroke is performed by a plurality of adiabatic compression strokes by a plurality of compressors and gas By replacing the sensible heat portion of the cooler and the final stage with a multi-stage isobaric heat release process using a condenser for cooling, it is possible to approximate an isothermal process (b → g).
By configuring the multi-stage cycle, the compression work Δ can be reduced and the COP (coefficient of performance) can be improved.

理論サイクル上の２つの等圧行程（行程ab及び行程cd）において、ガス冷媒と液冷媒の比熱が異なるため、ガス冷媒のエンタルピー差ΔＨab＝Ｈb−Ｈaと液冷媒のエンタルピー差ΔＨcd＝Ｈc−Ｈdが等しくならない。両者のエンタルピー差が等しくなる液冷媒の出口点を状態点ｈとするとΔＨab＝ΔＨchの関係が満足される。この液冷媒出口点ｈを「顕熱限界過冷却点」と定義する。なお、再生熱交換器の液冷媒出口温度が蒸発温度と等しくなる出口点ｄを「湿り限界過冷却点」と定義する。液出口点ｄまで液冷媒を過冷却させるためには、再生熱交換器のガス側吸入点を湿り状態点a’まで移動してΔＨa’b=ΔＨcdの関係を満足させる必要がある。ガス側吸入点をこのように移動するには第１膨張手段（例えば膨張弁）の開度を増加させて蒸発器への給液量を増加させれば良い。 In the two isobaric strokes in the theoretical cycle (stroke a and stroke cd), the specific heats of the gas refrigerant and the liquid refrigerant are different. Are not equal. If the exit point of the liquid refrigerant where the enthalpy difference between them is equal is the state point h, the relationship ΔHab = ΔHch is satisfied. This liquid refrigerant outlet point h is defined as “sensible heat limit supercooling point”. The outlet point d at which the liquid refrigerant outlet temperature of the regenerative heat exchanger becomes equal to the evaporation temperature is defined as the “wetness limit supercooling point”. In order to supercool the liquid refrigerant to the liquid outlet point d, it is necessary to move the gas side suction point of the regenerative heat exchanger to the wet state point a 'to satisfy the relationship ΔHa'b = ΔHcd. In order to move the gas side suction point in this way, the opening of the first expansion means (for example, an expansion valve) may be increased to increase the amount of liquid supplied to the evaporator.

図３及び図４は、単段圧縮行程および単段膨張行程を用いた逆エリクソンサイクルの再生熱交換器と成績係数の関係を示す。両図の縦軸は成績係数（ＣＯＰ）を示し、横軸は再生熱交換器の液冷媒出口温度（図３)およびガス冷媒出口温度（図４）を示す。グラフ線上の○、□、△、◇等の表示記号は理論計算点の位置を表示したものであり、実験結果を示すものではない。本特性は本発明の主要部を為すものであり、以下に記号Bの冷媒Ｒ６００ａ（イソブタン）を例にして若干詳しく説明する。 3 and 4 show the relationship between the regenerative heat exchanger and coefficient of performance of the reverse Ericsson cycle using the single-stage compression stroke and the single-stage expansion stroke. In both figures, the vertical axis represents the coefficient of performance (COP), and the horizontal axis represents the liquid refrigerant outlet temperature (FIG. 3) and the gas refrigerant outlet temperature (FIG. 4) of the regenerative heat exchanger. The symbols such as ○, □, Δ, and ◇ on the graph line indicate the positions of theoretical calculation points, and do not indicate experimental results. This characteristic is the main part of the present invention, and will be described in some detail below with reference to refrigerant R600a (isobutane) of symbol B as an example.

図３は再生熱交換器の液側冷媒の特性を冷媒別に示したものである。液冷媒出口温度４０℃線上の点ｃ（図１及び図２の状態点ｃ）は過冷却度が零、換言すれば再生熱交換器を使用しない状態のＣＯＰを示す。液冷媒出口温度−４０℃線上の点ｄ（図１及び図２の状態点ｄ）の性能は最大過冷却温度におけるＣＯＰを示す。図３より、再生熱交換器の液冷媒の出口温度が分れば直ちにそのときのＣＯＰを知ることができる。
液冷媒の出口温度はガス冷媒の入口温度及び湿り度により一意に決定される。例えば、ガス冷媒入口が飽和ガス状態であれば液冷媒出口は点ｈ（図１及び図２の状態点ｈ）となり、ＣＯＰは最大となる。そのとき液温度は図３より−１３℃程度である。この状態点ｈは前記の「顕熱限界過冷却点」である。 FIG. 3 shows the characteristics of the liquid side refrigerant of the regenerative heat exchanger for each refrigerant. The point c (state point c in FIGS. 1 and 2) on the liquid refrigerant outlet temperature 40 ° C. line indicates a COP in which the degree of supercooling is zero, in other words, the regenerative heat exchanger is not used. The performance of the point d (state point d in FIGS. 1 and 2) on the liquid refrigerant outlet temperature −40 ° C. line indicates the COP at the maximum supercooling temperature. From FIG. 3, if the outlet temperature of the liquid refrigerant in the regenerative heat exchanger is known, the COP at that time can be immediately known.
The outlet temperature of the liquid refrigerant is uniquely determined by the inlet temperature and the wetness of the gas refrigerant. For example, if the gas refrigerant inlet is in a saturated gas state, the liquid refrigerant outlet is at point h (state point h in FIGS. 1 and 2), and COP is maximized. At that time, the liquid temperature is about −13 ° C. from FIG. This state point h is the “sensible heat limit supercooling point” described above.

ガス冷媒の入口状態の湿り度を増加すると、液冷媒の過冷却温度を状態点ｄ（−４０℃）まで低下させることができる。状態点ｄが前記「湿り限界過冷却点」である。このときのガス冷媒の入口点（図１及び図２の状態点a’）を「湿り限界点a'」と定義する。ガス冷媒入口の湿り度を湿り限界点a'以上増加しても液冷媒出口温度は不変であり、ＣＯＰは逆に低下することになる。ガス冷媒入口温度が−４０℃以上の過熱温度になると、液冷媒出口状態が状態点ｈより高温状態となるため、図３の結果よりＣＯＰは低下する。 When the wetness of the inlet state of the gas refrigerant is increased, the supercooling temperature of the liquid refrigerant can be lowered to the state point d (−40 ° C.). The state point d is the “wet limit supercooling point”. The inlet point of the gas refrigerant at this time (state point a ′ in FIGS. 1 and 2) is defined as “wetting limit point a ′”. Even if the wetness degree of the gas refrigerant inlet is increased by the wetness limit point a ′ or more, the liquid refrigerant outlet temperature remains unchanged, and the COP decreases conversely. When the gas refrigerant inlet temperature reaches an overheating temperature of −40 ° C. or higher, the liquid refrigerant outlet state becomes a higher temperature state than the state point h, so that the COP decreases from the result of FIG.

図４は、再生熱交換器のガス側冷媒の特性を冷媒別に示したものである。
温度−４０℃線上のＣＯＰは再生熱交換器のガス冷媒入口が−４０℃の飽和ガス状態ａ（図１及び図２の状態点ａ）であり、かつガス冷媒出口が−４０℃の飽和ガス状態であるときのＣＯＰ、換言すれば、再生熱交換器を使用しない状態のＣＯＰを示す。ガス冷媒出口温度４０℃線上の点ｂ（図１及び図２の状態点ｂ）のＣＯＰは、液冷媒出口温度が図１及び図２の「顕熱限界過冷却点ｈ」となるときのＣＯＰを示す。
図４より、再生熱交換器のガス冷媒の出口温度が分れば、直ちにそのときのＣＯＰを知ることができる。ガス冷媒の出口温度が凝縮温度４０℃から低下するほどＣＯＰが低下する。 FIG. 4 shows the characteristics of the gas-side refrigerant of the regenerative heat exchanger for each refrigerant.
The COP on the temperature −40 ° C. line is a saturated gas state a (state point a in FIG. 1 and FIG. 2) at −40 ° C. at the gas refrigerant inlet of the regenerative heat exchanger and the saturated gas at the gas refrigerant outlet at −40 ° C. The COP in the state, that is, the COP in the state where the regenerative heat exchanger is not used is shown. The COP at the point b (state point b in FIGS. 1 and 2) on the gas refrigerant outlet temperature 40 ° C. line is the COP when the liquid refrigerant outlet temperature becomes the “sensible heat limit supercooling point h” in FIGS. Indicates.
From FIG. 4, if the outlet temperature of the gas refrigerant in the regenerative heat exchanger is known, the COP at that time can be immediately known. The COP decreases as the outlet temperature of the gas refrigerant decreases from the condensation temperature of 40 ° C.

図３で説明したように、「顕熱限界過冷却点ｈ」から「湿り限界過冷却点ｄ」の区間におけるＣＯＰは一定であることから、再生熱交換器による性能改善効果は、ガス冷媒の入口（飽和状態）と出口（過熱状態）の顕熱差を最大限利用したとき最大となることを意味する。
液冷媒の出口温度は、ガス冷媒の入口温度及び湿り度により一意的に決定されることはすでに述べた通りである。問題は、ガス冷媒の入口湿り度の影響が図４には一切示されていないことである。その理由は、ガス冷媒入口の湿り度が図１及び図２の状態ａから状態a’に移動しても入口温度は一定（蒸発温度）であるため、図４の座標上に湿り度の影響を示す方法が見当たらないことによる。
このことから、再生熱交換器のガス冷媒の入口を湿り状態に維持することは理論的性能改善に無関係であるが、ＣＯＰを最大化する制御法の観点から重要な意味をもつ。これについては後述する。 As described in FIG. 3, since the COP in the section from the “sensible heat limit supercooling point h” to the “wet limit supercooling point d” is constant, the performance improvement effect by the regenerative heat exchanger is It means that the maximum is obtained when the sensible heat difference between the inlet (saturated state) and the outlet (overheated state) is utilized to the maximum extent.
As described above, the outlet temperature of the liquid refrigerant is uniquely determined by the inlet temperature and the wetness of the gas refrigerant. The problem is that the effect of the inlet humidity of the gas refrigerant is not shown in FIG. The reason for this is that the inlet temperature is constant (evaporation temperature) even if the wetness of the gas refrigerant inlet moves from state a to state a ′ in FIGS. This is due to the lack of a way to indicate.
For this reason, maintaining the inlet of the gas refrigerant of the regenerative heat exchanger in a moist state is irrelevant to the theoretical performance improvement, but has an important meaning from the viewpoint of a control method for maximizing COP. This will be described later.

図３及び図４には次の二つの理論計算結果が示されている。
（１）再生熱交換器の冷端側および温端側の出入口端面における液側とガス側の冷媒温度差零℃の時の最大ＣＯＰ：液冷媒とガス冷媒の温度範囲はともに−４０℃から４０℃となり、ＣＯＰ最大値は図３の水平実線部分及び図４のガス冷媒出口温度４０℃におけるＣＯＰである。（図３及び図４の黒塗り点）
（２）再生熱交換器の冷端側および温端側の出入口端面における液側とガス側の冷媒温度差が５℃のときの最大ＣＯＰ：図３における−３５℃から４０℃及び図４における−４０℃から３５℃の温度範囲において、ＣＯＰ最大値は図３の水平破線部分及び図４のガス冷媒出口温度３５℃におけるＣＯＰである。（図３及び図４の白抜き点） 3 and 4 show the following two theoretical calculation results.
(1) Maximum COP when the refrigerant temperature difference between the liquid side and the gas side at the cold end side and hot end side inlet / outlet end faces of the regenerative heat exchanger is 0 ° C .: The temperature range of both the liquid refrigerant and the gas refrigerant is from −40 ° C. The COP maximum value is the COP at the horizontal solid line portion in FIG. 3 and the gas refrigerant outlet temperature of 40 ° C. in FIG. (Black dots in FIGS. 3 and 4)
(2) Maximum COP when the refrigerant temperature difference between the liquid side and the gas side at the inlet / outlet end surfaces on the cold end side and the warm end side of the regenerative heat exchanger is 5 ° C .: −35 ° C. to 40 ° C. in FIG. 3 and FIG. In the temperature range of −40 ° C. to 35 ° C., the COP maximum value is the COP at the horizontal broken line portion of FIG. 3 and the gas refrigerant outlet temperature of 35 ° C. of FIG. (Outline points in FIGS. 3 and 4)

前記（１）と（２）の二つの計算結果を示した理由は、再生熱交換器における冷端部及び温端部の温度差（例えば５℃）による伝熱損失が発生したとき、ＣＯＰがどの程度低下するかを確認するための計算値を示すためである。最大ＣＯＰを示す図３の水平部分の実線部（伝熱損失零）と水平破線部（伝熱損失を考慮）を比較することにより、伝熱損失の影響を実機のＣＯＰの変化として実感することができる。これらの計算結果を用いて正確に実機の運転条件におけるエリクソンサイクルのＣＯＰ特性を推定することが可能となる。以下にその理由を述べる。 The reason why the two calculation results (1) and (2) are shown is that when a heat transfer loss occurs due to a temperature difference (for example, 5 ° C.) between the cold end and the warm end in the regenerative heat exchanger, the COP is This is to show a calculated value for confirming how much it decreases. Realizing the effect of heat transfer loss as a change in actual COP by comparing the solid line part (zero heat transfer loss) and horizontal broken line part (considering heat transfer loss) in FIG. Can do. Using these calculation results, it is possible to accurately estimate the COP characteristic of the Ericsson cycle under the operating conditions of the actual machine. The reason is described below.

従来冷凍サイクルの理論成績係数をCOPtc（図３の４０℃における各冷媒のＣＯＰ、及び図４の−４０℃における各冷媒のＣＯＰ）、逆エリクソンサイクルの理論成績係数をCOPte(前記（１）の最大ＣＯＰ)、再生熱交換器の損失を考慮したときの理論成績係数をCOPteh（前記（２）の破線部ＣＯＰ）としたとき、同一運転条件における従来冷凍サイクルの実機成績係数をCOPpc、エリクソンサイクルの実機成績係数をCOPpeとする。エリクソンサイクルの実機成績係数COPpeは従来サイクルの実機成績係数COPpcを用いて次式より予測可能である。

ここで、式(c)のCOP_losは再生熱交換器損失によるＣＯＰ損失量を示す。 The theoretical coefficient of performance of the conventional refrigeration cycle is COPtc (the COP of each refrigerant at 40 ° C. in FIG. 3 and the COP of each refrigerant at −40 ° C. in FIG. 4), and the theoretical coefficient of performance of the reverse Ericsson cycle is COPte (the above (1) Maximum COP), and the theoretical coefficient of performance when considering the loss of the regenerative heat exchanger is COPteh (the broken line COP in (2) above), the actual coefficient of performance of the conventional refrigeration cycle under the same operating conditions is COPpc, Ericsson cycle Let COPpe be the actual coefficient of performance. The actual performance coefficient COPpe of the Ericsson cycle can be predicted from the following equation using the actual performance coefficient COPpc of the conventional cycle.

Here, COP _los in the equation (c) represents the amount of COP loss due to the loss of the regenerative heat exchanger.

式(a)より求めたエリクソンサイクルの実機成績係数が実験値と同程度に正確である理由は、式(b)の成績係数比Ｒ_copは理論計算値のため無誤差であり、式(c)における再生熱交換器損失を考慮したCOP_teh計算値は実用的に十分な精度で計算可能であることを考慮すると、COP_losの計算精度も十分であると考えられることによる。従って、従来サイクルの実測成績係数COP_pcの測定精度が十分であれば、これを用いてエリクソンサイクルの実機成績係数は式（ａ）より十分な精度で予測可能である。このことは、実用条件に合わせて実験確認をすることなく性能を十分な精度で予測できることを意味する。さらに、以下に示すように、再生熱交換器損失による成績係数への影響は理論成績係数の１%以下の程度であることを考慮すると、式（ａ）による実機性能予測値は誤差１%を超えることはあり得ず、実験結果の精度範囲にあると考えることができる。 The reason why the actual coefficient of performance of the Ericsson cycle calculated from the equation (a) is as accurate as the experimental value is that the coefficient of performance ratio R _cop of the equation (b) is a theoretical calculation value and is error-free. This is because the calculation accuracy of COP _los is considered to be sufficient considering that the calculated COP _teh value considering the regenerative heat exchanger loss in) can be calculated with sufficient accuracy. Therefore, if the measurement accuracy of the actual performance coefficient COP _pc of the conventional cycle is sufficient, the actual performance coefficient of the Ericsson cycle can be predicted with sufficient accuracy using the equation (a). This means that performance can be predicted with sufficient accuracy without experimental confirmation according to practical conditions. Furthermore, as shown below, considering that the effect on the coefficient of performance due to the loss of the regenerative heat exchanger is about 1% or less of the theoretical coefficient of performance, the actual machine performance prediction value by the formula (a) has an error of 1%. It cannot be exceeded and can be considered to be within the accuracy range of the experimental results.

図３及び図４に例示された冷媒のＣＯＰ損失量COP_losを計算すると、(−０．０１０＜COP_los＜０．０２５)程度であり、理論成績係数COP_teに対する損失Ｒ_los（＝COP_los／COP_teは、−０．５８％（Ｒ７１７）＜Ｒ_los＜１．１％（Ｒ６００ａ）となる。 The COP loss amount COP _los of the refrigerant exemplified in FIGS. 3 and 4 is calculated to be about (−0.010 <COP _los <0.025), and the loss R _los (= COP _{los with} respect to the theoretical coefficient of performance COP _te / COP _te is −0.58% (R717) <R _los <1.1% (R600a).

例えば、図３及び図４中、冷媒Ｒ６００ａ（イソブタン）の実線成績係数（ＣＯＰ）は、再生熱交換器の伝熱損失が零の場合の性能を示す。液冷媒出口温度が４０℃（過冷却度０℃）でＣＯＰ＝１．９２であり、顕熱限界過冷却点ｈにおける顕熱限界過冷却温度ＴｈとＣＯＰはTh=−１３℃、ＣＯＰ＝２．３０となる。過冷却温度が−４０℃となると、湿り限界過冷却点ｄ（温度Td）における温度とＣＯＰはそれぞれTd=−４０℃、ＣＯＰ=２．３０である。
このことは、再生熱交換器の液冷媒出口温度が限界過冷却温度Tdと顕熱限界過冷却温度Thの温度範囲内であるとき成績係数が最大かつ一定となることを示している。 For example, in FIGS. 3 and 4, the solid line coefficient of performance (COP) of the refrigerant R600a (isobutane) indicates the performance when the heat transfer loss of the regenerative heat exchanger is zero. The liquid refrigerant outlet temperature is 40 ° C. (supercooling degree 0 ° C.), COP = 1.92, the sensible heat limit supercooling temperature Th and COP at the sensible heat limit supercooling point h are Th = −13 ° C., COP = 2 .30. When the supercooling temperature is −40 ° C., the temperature and COP at the wetness limit supercooling point d (temperature Td) are Td = −40 ° C. and COP = 2.30, respectively.
This indicates that the coefficient of performance becomes maximum and constant when the liquid refrigerant outlet temperature of the regenerative heat exchanger is within the temperature range of the limit supercooling temperature Td and the sensible heat limit subcooling temperature Th.

同様に、再生熱交換器の伝熱損失を考慮して、冷端側および温端側出入り口の液冷媒とガス冷媒の温度差を５℃としたときの性能を水平破線で示す。破線部以外の性能は伝熱損失の有無に無関係に同一性能となるため一本の実線で表示されている。
図３において、液側冷媒の温度範囲は−３５℃から４０℃であり、図４においてガス側冷媒の温度範囲は−４０℃から３５℃である。再生熱交換器の冷端側温度差が５℃であることから、図４の液冷媒出口温度は−３５℃であり、温端温度差が５℃であるガス冷媒出口温度は３５℃である。図３より明らかなように、再生熱交換器の伝熱損失の増加に合わせて「顕熱限界過冷却点」は高温側に移動して最大ＣＯＰは低下していく。 Similarly, in view of the heat transfer loss of the regenerative heat exchanger, the performance when the temperature difference between the liquid refrigerant and the gas refrigerant at the cold end side and the hot end side inlet / outlet is 5 ° C. is indicated by a horizontal broken line. Since the performance other than the broken line portion is the same regardless of the presence or absence of heat transfer loss, it is indicated by a single solid line.
In FIG. 3, the temperature range of the liquid side refrigerant is −35 ° C. to 40 ° C., and in FIG. 4, the temperature range of the gas side refrigerant is −40 ° C. to 35 ° C. Since the cold end side temperature difference of the regenerative heat exchanger is 5 ° C., the liquid refrigerant outlet temperature in FIG. 4 is −35 ° C., and the gas refrigerant outlet temperature where the hot end temperature difference is 5 ° C. is 35 ° C. . As is clear from FIG. 3, the “sensible heat limit supercooling point” moves to the high temperature side and the maximum COP decreases as the heat transfer loss of the regenerative heat exchanger increases.

なお、Ｒ７１７（ＮＨ_３）の場合は、液側過冷却による冷凍能力の増加よりもガス側で圧縮機吐出温度の上昇による圧縮仕事Δの増加の影響のほうが大きいためＣＯＰの傾向が逆になる。但し、圧縮機の段数を増加すれば各圧縮機の吐出温度を低下でき圧縮損失が改善できるので、ＣＯＰを向上できる。 In the case of R717 (NH ₃ ), the influence of the increase in the compression work Δ due to the increase in the compressor discharge temperature on the gas side is greater than the increase in the refrigerating capacity due to the liquid side subcooling, so the tendency of COP is reversed. . However, if the number of stages of the compressor is increased, the discharge temperature of each compressor can be lowered and the compression loss can be improved, so that the COP can be improved.

冷媒Ａ（Ｒ７１７：アンモニア）は、単段圧縮行程および単段膨張行程を持つ逆エリクソンサイクルにおいて他の例示冷媒と著しく異なるサイクル性能を持つ。即ち、再生熱交換器による液冷媒の過冷却により性能は急速に低下する。本発明の主旨からはずれるため説明は省略するが、圧縮段数および膨張段数を増加することにより、冷媒Ｒ７１７といえども冷媒Ｒ６００ａと同程度に性能を改善させることが可能である。 Refrigerant A (R717: ammonia) has a cycle performance that is significantly different from other exemplary refrigerants in a reverse Ericsson cycle having a single stage compression stroke and a single stage expansion stroke. That is, the performance rapidly decreases due to the supercooling of the liquid refrigerant by the regenerative heat exchanger. Although explanation is omitted because it deviates from the gist of the present invention, by increasing the number of compression stages and the number of expansion stages, even the refrigerant R717 can improve the performance to the same extent as the refrigerant R600a.

幾つかの実施形態では、前記構成（１）において、
（２）前記凝縮器における冷媒の凝縮圧力をＰｃとし、前記蒸発器における冷媒の蒸発圧力をＰｅとし、前記複数段の圧縮機の段数をｎとしたとき、前記複数段の断熱圧縮行程において前記複数段の圧縮機間の圧縮比ｒを（Ｐｃ／Ｐｅ）^１／ｎとし、
前記制御装置は、前記ガス冷却器による前記複数段の等圧放熱行程において、入口ガス冷媒（凝縮温度＋α）を凝縮温度近傍まで冷却するものである。ここでαは各段圧縮機の吸入ガスの圧縮による昇温幅であり、段数を増やすことによりαは低下させることができる。このため、圧縮機の機械的な使用温度範囲や多段化による性能改善を考慮して、αを選定するのがよい。 In some embodiments, in the configuration (1),
(2) When the condensing pressure of the refrigerant in the condenser is Pc, the evaporating pressure of the refrigerant in the evaporator is Pe, and the number of stages of the plurality of compressors is n, the adiabatic compression stroke in the plurality of stages The compression ratio r between the multi-stage compressors is (Pc / Pe) ^{1 / n} ,
The control device cools the inlet gas refrigerant (condensation temperature + α) to the vicinity of the condensation temperature in the multi-stage isobaric heat radiation process by the gas cooler. Here, α is the range of temperature rise due to the compression of the intake gas of each stage compressor, and α can be lowered by increasing the number of stages. For this reason, it is preferable to select α in consideration of the mechanical operating temperature range of the compressor and performance improvement due to multistage.

前記構成（２）によれば、前記複数段の断熱圧縮行程において、前記複数段の圧縮機間の圧縮比ｒを（Ｐｃ／Ｐｅ）^１／ｎとしたことで、前記複数段の圧縮機におけるガス冷媒の吐出圧力及び吐出温度をほぼ同等にでき、一部の圧縮機で吐出圧力及び吐出温度が突出するのを抑えることができる。これによって、潤滑油の劣化やパッキン材などの焼損を抑制できる。 According to the configuration (2), in the multiple-stage adiabatic compression process, the compression ratio r between the multiple-stage compressors is (Pc / Pe) ^{1 / n} . The discharge pressure and discharge temperature of the gas refrigerant can be made substantially equal, and the discharge pressure and discharge temperature can be prevented from protruding in some compressors. As a result, deterioration of the lubricating oil and burning of the packing material can be suppressed.

幾つかの実施形態では、前記構成（１）又は（２）において、
（３）前記再生熱交換器と前記第１膨張手段との間の前記冷媒循環路に設けられた第１液ガス分離器と、
前記第１液ガス分離器の入口側の前記冷媒循環路に設けられた第２膨張手段と、
前記第１液ガス分離器のガス冷媒を前記複数段の圧縮機のうちの低段側圧縮機の出口と次段の圧縮機の入口に接続された冷媒路に供給する中間ガスラインと、を有する中間冷却装置をさらに備えている。
前記構成（３）によれば、前記中間冷却装置を設けることでエンタルピ差が増加し、それに対応した全圧縮仕事の変化が起こるが、結果としてＣＯＰが向上する。従って、冷凍又は空調装置のＣＯＰをさらに向上できる。 In some embodiments, in the configuration (1) or (2),
(3) a first liquid gas separator provided in the refrigerant circuit between the regeneration heat exchanger and the first expansion means;
A second expansion means provided in the refrigerant circuit on the inlet side of the first liquid gas separator;
An intermediate gas line for supplying a gas refrigerant of the first liquid gas separator to a refrigerant path connected to an outlet of a low-stage compressor of the plurality of stages of compressors and an inlet of a next-stage compressor; And an intermediate cooling device.
According to the configuration (3), the enthalpy difference is increased by providing the intermediate cooling device, and the corresponding total compression work is changed, but as a result, the COP is improved. Therefore, the COP of the refrigeration or air conditioner can be further improved.

幾つかの実施形態では、前記構成（１）〜（３）の何れかにおいて、
（４）前記再生熱交換器と前記第１膨張手段との間の前記冷媒循環路に設けられた第２液ガス分離器と、
前記第２液ガス分離器の入口側の前記冷媒循環路に設けられた第３膨張手段と、
前記第２液ガス分離器のガス冷媒を前記複数段の圧縮機のうちの低段側圧縮機の中間圧領域に供給するエコノマイザガスラインと、を有するエコノマイザ装置をさらに備えている。
前記構成（４）によれば、前記エコノマイザ装置を設けることで、エンタルピ差が増加し、それに対応した全圧縮仕事の変化が起こるが、結果としてＣＯＰが向上する。従って、冷凍又は空調装置のＣＯＰをさらに向上できる。 In some embodiments, in any of the configurations (1) to (3),
(4) a second liquid gas separator provided in the refrigerant circuit between the regenerative heat exchanger and the first expansion means;
Third expansion means provided in the refrigerant circuit on the inlet side of the second liquid gas separator;
And an economizer gas line for supplying the gas refrigerant of the second liquid gas separator to an intermediate pressure region of a low stage compressor of the plurality of stages of compressors.
According to the configuration (4), by providing the economizer device, the difference in enthalpy increases and a corresponding change in total compression work occurs, but as a result, COP is improved. Therefore, the COP of the refrigeration or air conditioner can be further improved.

幾つかの実施形態では、前記構成（１）〜（４）の何れかにおいて、
（５）前記複数段の圧縮機の段数が２段又は３段である。
前記複数段の圧縮機の段数が４以上になると、設備費の増加の割にＣＯＰ向上に寄与しないので、２段又は３段が実用的な段数となる。 In some embodiments, in any of the configurations (1) to (4),
(5) The number of stages of the plurality of stages of compressors is two or three.
If the number of stages of the plurality of compressors is 4 or more, it does not contribute to the improvement of COP for the increase in equipment cost, so 2 or 3 stages are practical stages.

本発明の少なくとも一実施形態に係る冷凍又は空調装置の制御方法は、
（６）冷媒循環路に直列に設けられた複数段の圧縮機、凝縮器、第１膨張手段及び蒸発器と、
前記複数段の圧縮機で各圧縮機の出口と次段の圧縮機の入口と結ぶ前記冷媒循環路に設けられ、各圧縮機の吐出ガス冷媒を冷却するためのガス冷却器と、
前記蒸発器より前記圧縮機に向かうガス冷媒と前記凝縮器から前記第１膨張手段に向かう液冷媒とを熱交換させるための再生熱交換器と、を備えた冷凍又は空調装置の制御方法において、
主として前記凝縮器による等温放熱行程と、前記蒸発器による等温吸熱行程と、前記再生熱交換器における熱交換によって行われる液領域における等圧放熱行程及び過熱蒸気領域における等圧吸熱行程を含む逆エリクソンサイクルを形成すると共に、
前記等温放熱行程のうち過熱蒸気領域で行われる部分行程が、前記複数段の圧縮機による複数段の断熱圧縮行程と、前記ガス冷却器及び前記凝縮器による複数段の等圧放熱行程とに置き換えた第１行程と、
前記等圧吸熱行程で湿り状態の冷媒を前記再生熱交換器に導入し、前記等圧吸熱行程で飽和ガス冷媒が前記再生熱交換器に導入されたときの前記等温放熱行程の終了点となる顕熱限界過冷却点から、前記第１膨張手段による断熱膨張行程の終了点が飽和冷媒液線上に位置するときの前記等温放熱行程の終了点となる湿り限界過冷却点までの領域の過冷却点に前記等温放熱行程の終了点が位置するように制御する第２行程と、を含んでいる。 A control method for a refrigeration or air-conditioning apparatus according to at least one embodiment of the present invention includes:
(6) a plurality of stages of compressors, condensers, first expansion means and evaporators provided in series in the refrigerant circuit;
A gas cooler for cooling the discharged gas refrigerant of each compressor, provided in the refrigerant circulation path connecting the outlet of each compressor and the inlet of the next stage compressor in the plurality of stages of compressors;
In a control method of a refrigeration or air conditioner comprising a regenerative heat exchanger for exchanging heat between a gas refrigerant from the evaporator toward the compressor and a liquid refrigerant from the condenser toward the first expansion means,
Inverse Ericsson mainly including an isothermal heat release process in the condenser, an isothermal heat absorption process in the evaporator, an isobaric heat release process in the liquid region and an isobaric heat absorption process in the superheated steam region performed by heat exchange in the regenerative heat exchanger Forming a cycle,
The partial process performed in the superheated steam region in the isothermal heat dissipation process is replaced with a plurality of adiabatic compression processes by the multistage compressor and a multistage isobaric heat dissipation process by the gas cooler and the condenser. The first step,
The wet refrigerant is introduced into the regenerative heat exchanger in the isobaric endothermic process, and is the end point of the isothermal heat releasing process when the saturated gas refrigerant is introduced into the regenerative heat exchanger in the isobaric endothermic process. Supercooling of the region from the sensible heat limit supercooling point to the wetness limit supercooling point which is the end point of the isothermal heat release process when the end point of the adiabatic expansion process by the first expansion means is located on the saturated refrigerant liquid line And a second process for controlling the end point of the isothermal heat radiation process to be located at a point.

前記構成（６）によれば、前記第１行程により、実用サイクルを理論上の逆エリクソンサイクルに近づけることができるので、ＣＯＰを向上できると共に、ガス冷媒の温度上昇を抑制できるため、冷媒に含まれる潤滑油の劣化を防止できると共に、パッキン材などを構成するエラストマの焼損を防止できる。
さらに、前記第２行程により、冷凍又は空調装置のＣＯＰを最大に維持できると共に、最適制御条件が点設定でなく範囲設定となるため制御が極めて容易である。 According to the configuration (6), since the practical cycle can be brought close to the theoretical inverse Ericsson cycle by the first stroke, the COP can be improved and the temperature rise of the gas refrigerant can be suppressed. It is possible to prevent deterioration of the lubricating oil, and to prevent burning of the elastomer constituting the packing material.
Further, the second process can maintain the maximum COP of the refrigeration or air conditioner, and the control is extremely easy because the optimum control condition is not a point setting but a range setting.

幾つかの実施形態では、前記構成（６）において、
（７）前記第２行程は、少なくとも前記複数段の圧縮機を駆動するモータの回転数制御と、前記第１膨張手段の制御とで行われる。
前記構成（７）によれば、従来から行われている簡単な制御でＣＯＰを向上できる。 In some embodiments, in the configuration (6),
(7) The second stroke is performed by at least the rotation speed control of a motor that drives the plurality of stages of compressors and the control of the first expansion means.
According to the configuration (7), COP can be improved by simple control that has been conventionally performed.

（８）幾つかの実施形態では、前記構成（６）又は（７）において、
前記第２行程は、
前記凝縮器から前記第１膨張手段に向かう液冷媒の前記再生熱交換器出口における温度を検出する第１ステップと、
前記第１ステップで検出した温度検出値を、前記顕熱限界過冷却点における冷媒温度から前記蒸発器における冷媒の蒸発温度までの間の温度に制御する第２ステップと、を含んでいる。
前記構成（８）によれば、前記温度検出値は蒸発器や再生熱交換器に設けた温度センサの検出値から、顕熱限界過冷却点や湿り限界過冷却点を容易に求めることができる。 (8) In some embodiments, in the configuration (6) or (7),
The second stroke is
A first step of detecting a temperature at the outlet of the regenerative heat exchanger of liquid refrigerant from the condenser toward the first expansion means;
And a second step of controlling the temperature detection value detected in the first step to a temperature between the refrigerant temperature at the sensible heat limit supercooling point and the refrigerant evaporation temperature in the evaporator.
According to the configuration (8), the sensible heat limit subcooling point and the wetness limit subcooling point can be easily obtained from the temperature detection value from the detection value of the temperature sensor provided in the evaporator or the regenerative heat exchanger. .

本発明の少なくとも一実施形態によれば、近似逆エリクソンサイクルを用いた冷凍又は空調装置において、熱効率のさらなる向上が可能になる。 According to at least one embodiment of the present invention, it is possible to further improve thermal efficiency in a refrigeration or air conditioner using an approximate inverse Ericsson cycle.

近似逆エリクソンサイクルのＴ−Ｓ線図である。It is a TS diagram of an approximate reverse Ericsson cycle. 近似逆エリクソンサイクルのモリエル線図である。It is a Mollier diagram of an approximate inverse Ericsson cycle. 再生熱交換器の液冷媒出口温度とＣＯＰの関係を示す線図である。It is a diagram which shows the relationship between the liquid refrigerant exit temperature of a regeneration heat exchanger, and COP. 再生熱交換器のガス冷媒出口温度とＣＯＰの関係を示す線図である。It is a diagram which shows the relationship between the gas refrigerant exit temperature of a regenerative heat exchanger, and COP. 一実施形態に係る冷凍装置の系統図である。It is a systematic diagram of the freezing apparatus concerning one embodiment. 図５に示す冷凍装置の近似逆エリクソンサイクルを示すＴ−Ｓ線図である。FIG. 6 is a TS diagram showing an approximate inverse Ericsson cycle of the refrigeration apparatus shown in FIG. 5. 一実施形態に係る冷凍装置の系統図である。It is a systematic diagram of the freezing apparatus concerning one embodiment. 図７に示す冷凍装置の近似逆エリクソンサイクルを示すＴ−Ｓ線図である。It is a TS diagram which shows the approximate reverse Ericsson cycle of the freezing apparatus shown in FIG. 図７に示す冷凍装置の近似逆エリクソンサイクルを示すモリエル線図である。FIG. 8 is a Mollier diagram showing an approximate inverse Ericsson cycle of the refrigeration apparatus shown in FIG. 7. 一実施形態に係る冷凍装置の系統図である。It is a systematic diagram of the freezing apparatus concerning one embodiment. 図１０に示す冷凍装置の近似逆エリクソンサイクルを示すＴ−Ｓ線図である。It is a TS diagram which shows the approximate reverse Ericsson cycle of the freezing apparatus shown in FIG. 一実施形態に係る冷凍装置の系統図である。It is a systematic diagram of the freezing apparatus concerning one embodiment. 一実施形態に係る冷凍装置の系統図である。It is a systematic diagram of the freezing apparatus concerning one embodiment. 図１３に示す冷凍装置が構成する近似逆エリクソンサイクルを示すＴ−Ｓ線図である。It is a TS diagram which shows the approximate reverse Ericsson cycle which the refrigeration apparatus shown in FIG. 13 comprises. 前記実施形態に係る冷凍装置のＣＯＰ（計算値）を示す図表である。It is a graph which shows COP (calculated value) of the freezing apparatus which concerns on the said embodiment. 近似逆エリクソンサイクル及び従来冷凍サイクルのＣＯＰを示す線図である。It is a diagram which shows COP of an approximate reverse Ericsson cycle and a conventional refrigerating cycle.

以下、添付図面を参照して本発明の幾つかの実施形態について説明する。ただし、実施形態として記載され又は図面に示されている構成部品の寸法、材質、形状、その相対的配置等は、本発明の範囲をこれに限定する趣旨ではなく、単なる説明例にすぎない。
例えば、「ある方向に」、「ある方向に沿って」、「平行」、「直交」、「中心」、「同心」或いは「同軸」等の相対的或いは絶対的な配置を表す表現は、厳密にそのような配置を表すのみならず、公差、若しくは、同じ機能が得られる程度の角度や距離をもって相対的に変位している状態も表すものとする。
例えば、「同一」、「等しい」及び「均質」等の物事が等しい状態であることを表す表現は、厳密に等しい状態を表すのみならず、公差、若しくは、同じ機能が得られる程度の差が存在している状態も表すものとする。
例えば、四角形状や円筒形状等の形状を表す表現は、幾何学的に厳密な意味での四角形状や円筒形状等の形状を表すのみならず、同じ効果が得られる範囲で、凹凸部や面取り部等を含む形状も表すものとする。
一方、一つの構成要素を「備える」、「具える」、「具備する」、「含む」、又は「有する」という表現は、他の構成要素の存在を除外する排他的な表現ではない。 Hereinafter, some embodiments of the present invention will be described with reference to the accompanying drawings. However, the dimensions, materials, shapes, relative arrangements, and the like of the components described in the embodiments or shown in the drawings are not intended to limit the scope of the present invention, but are merely illustrative examples.
For example, expressions expressing relative or absolute arrangements such as “in a certain direction”, “along a certain direction”, “parallel”, “orthogonal”, “center”, “concentric” or “coaxial” are strictly In addition to such an arrangement, it is also possible to represent a state of relative displacement with an angle or a distance such that tolerance or the same function can be obtained.
For example, an expression indicating that things such as “identical”, “equal”, and “homogeneous” are in an equal state not only represents an exactly equal state, but also has a tolerance or a difference that can provide the same function. It also represents the existing state.
For example, expressions representing shapes such as quadrangular shapes and cylindrical shapes represent not only geometrically strict shapes such as quadrangular shapes and cylindrical shapes, but also irregularities and chamfers as long as the same effects can be obtained. A shape including a part or the like is also expressed.
On the other hand, the expressions “comprising”, “comprising”, “comprising”, “including”, or “having” one constituent element are not exclusive expressions for excluding the existence of other constituent elements.

（実施形態１）
図５〜図１２は、本発明に係る冷凍又は空調装置の幾つかの実施形態を示している。このうち、図５は３段圧縮３段膨張を行う冷凍装置１０Ａを示し、図６は冷凍装置１０Ａが構成する近似逆エリクソンサイクルのＴ−Ｓ線図である。
図７は３段圧縮２段膨張を行う冷凍装置１０Ｂを示し、図８は冷凍装置１０Ｂが構成する近似逆エリクソンサイクルのＴ−Ｓ線図であり、図９は該近似逆エリクソンサイクルのモリエル線図である。
図１０は３段圧縮単段膨張を行う冷凍装置１０Ｃを示し、図１１は冷凍装置１０Ｃが構成する近似逆エリクソンサイクルのＴ−Ｓ線図である。
図１２は３段圧縮２段膨張を行う冷凍装置１０Ｄを示している。 (Embodiment 1)
5 to 12 show some embodiments of the refrigeration or air conditioning apparatus according to the present invention. 5 shows a refrigeration apparatus 10A that performs three-stage compression and three-stage expansion, and FIG. 6 is a TS diagram of the approximate inverse Ericsson cycle that the refrigeration apparatus 10A configures.
FIG. 7 shows a refrigeration apparatus 10B that performs three-stage compression and two-stage expansion, FIG. 8 is a TS diagram of an approximate inverse Ericsson cycle constituted by the refrigeration apparatus 10B, and FIG. 9 is a Mollier line of the approximate inverse Ericsson cycle. FIG.
FIG. 10 shows a refrigeration apparatus 10C that performs three-stage compression single-stage expansion, and FIG. 11 is a TS diagram of an approximate inverse Ericsson cycle that is constituted by the refrigeration apparatus 10C.
FIG. 12 shows a refrigeration apparatus 10D that performs three-stage compression and two-stage expansion.

図５〜図１２に示す冷凍装置１０Ａ〜１０Ｄは、冷媒循環路１２に冷凍サイクルを構成する複数段（３段）の圧縮機１３ａ、１３ｂ及び１３ｃ、凝縮器１６、第１膨張弁１８及び蒸発器２０が直列に設けられている。各圧縮機は夫々電動モータ１４ａ、１４ｂ及び１４ｃによって回転駆動される。前記圧縮機は、例えばスクリュー圧縮機などの容積型圧縮機が用いられる。但し、運転条件の変動が少ない用途にターボ式圧縮機が使われることもある。
各圧縮機の出口と入口とを接続するガス冷媒路１２ａ及び１２ｂに、冷却水などの冷却媒体で圧縮機１３ａ又は１３ｂから吐出されたガス冷媒を冷却するガス冷却器２２ａ及び２２ｂを備えている。
さらに、蒸発器２０より圧縮機１３ａに向かうガス冷媒と凝縮器１６から出た液冷媒とを熱交換させるための再生熱交換器２４を備えている。 The refrigeration apparatuses 10A to 10D shown in FIGS. 5 to 12 include a plurality of stages (three stages) of compressors 13a, 13b and 13c, a condenser 16, a first expansion valve 18 and evaporation constituting the refrigeration cycle in the refrigerant circulation path 12. A vessel 20 is provided in series. Each compressor is rotationally driven by electric motors 14a, 14b and 14c, respectively. As the compressor, for example, a positive displacement compressor such as a screw compressor is used. However, a turbo compressor may be used for an application in which fluctuations in operating conditions are small.
The gas refrigerant paths 12a and 12b connecting the outlet and the inlet of each compressor are provided with gas coolers 22a and 22b for cooling the gas refrigerant discharged from the compressor 13a or 13b with a cooling medium such as cooling water. .
Furthermore, a regenerative heat exchanger 24 for exchanging heat between the gas refrigerant headed from the evaporator 20 toward the compressor 13a and the liquid refrigerant discharged from the condenser 16 is provided.

冷凍装置１０Ａ及び１０Ｂの例示的な構成として、再生熱交換器２４と第１膨張弁１８との間の冷媒循環路１２に設けられた第１液ガス分離器２８と、第１液ガス分離器２８の入口側の冷媒循環路１２に設けられた第２膨張弁３０と、第１液ガス分離器２８で液冷媒と分離されたガス冷媒を第１段圧縮機１３ａの出口と第２段圧縮機１３ｂの入口とに接続された冷媒路１２ａに供給する中間ガス路３２とを有する中間冷却装置２６をさらに備えている。中間ガス路３２はガス冷却器２２ａの下流側で冷媒路１２ａに接続されている。 As an exemplary configuration of the refrigeration apparatuses 10A and 10B, a first liquid gas separator 28 provided in the refrigerant circulation path 12 between the regenerative heat exchanger 24 and the first expansion valve 18, and a first liquid gas separator 28, the second expansion valve 30 provided in the refrigerant circulation path 12 on the inlet side, and the gas refrigerant separated from the liquid refrigerant by the first liquid gas separator 28, and the second stage compression of the gas refrigerant separated from the liquid refrigerant. It further includes an intermediate cooling device 26 having an intermediate gas passage 32 that supplies the refrigerant passage 12a connected to the inlet of the machine 13b. The intermediate gas path 32 is connected to the refrigerant path 12a on the downstream side of the gas cooler 22a.

冷凍装置１０Ａ及び１０Ｄの例示的な構成として、再生熱交換器２４と第１膨張弁１８との間の冷媒循環路１２に設けられた第２液ガス分離器３６と、第２液ガス分離器３６の入口側の冷媒循環路１２に設けられた第３膨張弁３８と、第２液ガス分離器３６のガス冷媒を第１段圧縮機１３ａの中間圧領域に供給するエコノマイザガス路４０とを有するエコノマイザ装置３４をさらに備えている。
なお、冷凍装置１０Ａでは、第１液ガス分離器２８の液冷媒を冷媒循環路１２を通して第２液ガス分離器３６に供給し、冷凍装置１０Ｂでは再生熱交換器２４から冷媒循環路１２を介して第１液ガス分離器２８に液冷媒を供給している。 As an exemplary configuration of the refrigeration apparatuses 10A and 10D, a second liquid gas separator 36 provided in the refrigerant circuit 12 between the regenerative heat exchanger 24 and the first expansion valve 18, and a second liquid gas separator A third expansion valve 38 provided in the refrigerant circulation path 12 on the inlet side of 36, and an economizer gas path 40 for supplying the gas refrigerant of the second liquid gas separator 36 to the intermediate pressure region of the first stage compressor 13a. An economizer device 34 is further provided.
In the refrigeration apparatus 10A, the liquid refrigerant in the first liquid gas separator 28 is supplied to the second liquid gas separator 36 through the refrigerant circuit 12, and in the refrigeration apparatus 10B, the regenerative heat exchanger 24 passes through the refrigerant circuit 12. The liquid refrigerant is supplied to the first liquid gas separator 28.

図中、符号Ｐは圧力センサの配置を示し、Ｔは温度センサの配置を示し、Ｇは冷媒液面のレベルセンサの配置を示している。冷凍装置１０Ａ〜１０Ｄの各所に設けられた前記温度センサ、前記圧力センサ及び前記レベルセンサ、駆動モータ１４ａ、１４ｂ及び１４ｃに設けられた回転数センサ４４ａ、４４ｂ及び４４ｃ等の検出値はすべて制御装置４２に入力される。
制御装置４２は、これらの検出値に基づいて、第１膨張弁１８、第２膨張弁３０及び第３膨張弁３８の開度、及び電動モータ１４ａ、１４ｂ及び１４ｃの回転数、中間ガスライン３２に設けられた流量調整弁４６の開度及びエコノマイザガス路４０に設けられた流量調整弁４８の開度等を制御する。 In the figure, symbol P indicates the arrangement of the pressure sensor, symbol T indicates the arrangement of the temperature sensor, and symbol G indicates the arrangement of the level sensor on the refrigerant liquid level. All the detected values of the temperature sensors, the pressure sensors, the level sensors, and the rotation speed sensors 44a, 44b, and 44c provided in the drive motors 14a, 14b, and 14c are provided in various places in the refrigeration apparatuses 10A to 10D. 42.
Based on these detection values, the control device 42 opens the first expansion valve 18, the second expansion valve 30 and the third expansion valve 38, the rotational speeds of the electric motors 14 a, 14 b and 14 c, and the intermediate gas line 32. And the opening degree of the flow rate adjusting valve 48 provided in the economizer gas path 40 is controlled.

かかる構成において、冷凍装置１０Ａ〜１０Ｄは理論上の逆エリクソンサイクルに近似した前述の実用的逆エリクソンサイクルを構成する。冷凍装置１０Ａ〜１０Ｄが構成する近似逆エリクソンサイクルは、図６、図８及び図１１のＴ−Ｓ線図、及び図９のモリエル線図に示される。
冷凍装置１０Ａ〜１０Ｄの近似逆エリクソンサイクルは、主として凝縮器１６による等温放熱行程（ｂ→ｃ）と、蒸発器２０による等温吸熱行程（ｆ→ａ）と、再生熱交換器２４によって行われる液領域の等圧放熱行程（ｃ→ｈ）及び過熱蒸気領域の等圧吸熱行程（ａ→ｂ）を含んでいる。
また、等温放熱行程（ｂ→ｃ）のうち過熱蒸気領域で行われる部分行程が、複数段の圧縮機１３ａ、１３ｂ及び１３ｃによる複数段の断熱圧縮行程（ｂ→ｂ’）と、ガス冷却器２２ａ及び２２ｂ及び凝縮器１６による複数段の等圧放熱行程（ｂ’→ｂ）及び凝縮器１６による１段の等圧放熱行程（ｂ’→ｇ）とに置き換えられる（第１行程）。 In such a configuration, the refrigeration apparatuses 10A to 10D constitute the practical inverse Ericsson cycle described above that approximates the theoretical inverse Ericsson cycle. The approximate inverse Ericsson cycle constituted by the refrigeration apparatuses 10A to 10D is shown in the TS diagrams of FIGS. 6, 8 and 11, and the Mollier diagram of FIG.
The approximate inverse Ericsson cycle of the refrigeration apparatuses 10A to 10D mainly includes an isothermal heat release process (b → c) by the condenser 16, an isothermal endothermic process (f → a) by the evaporator 20, and a liquid performed by the regenerative heat exchanger 24. It includes the isobaric heat release process (c → h) of the region and the isobaric heat absorption process (a → b) of the superheated steam region.
Moreover, the partial process performed in a superheated steam area | region among isothermal heat radiation processes (b-> c), the multistage adiabatic compression process (b-> b ') by the multistage compressors 13a, 13b, and 13c, and a gas cooler It is replaced with a multi-stage isobaric heat release process (b ′ → b) by 22a and 22b and the condenser 16 and a single-stage isobaric heat release process (b ′ → g) by the condenser 16 (first process).

冷凍装置１０Ａ〜１０Ｄの例示的な構成では、凝縮器１６における凝縮圧力をＰｃ、蒸発器２０における蒸発圧力をＰｅ、複数段の圧縮機の段数をｎとしたとき、各圧縮機間の圧縮比を（Ｐｃ／Ｐｅ）^１／ｎとする。冷凍装置１０Ａ〜１０Ｄは３段の圧縮機１３ａ、１３ｂ及び１３ｃを有するので、各圧縮機間の圧縮比は（Ｐｃ／Ｐｅ）^１／３となる。
また、ガス冷却器２２ａ及び２２ｂによる２段の等圧放熱行程において、吐出ガス冷媒温度を（凝縮温度＋α）として、各段の等圧放熱工程でα分を冷却することで、一部の圧縮機での吐出温度の突出を抑制して均質な圧縮と放熱を実現する。 In the exemplary configurations of the refrigeration apparatuses 10A to 10D, when the condensation pressure in the condenser 16 is Pc, the evaporation pressure in the evaporator 20 is Pe, and the number of stages of the plurality of compressors is n, the compression ratio between the compressors Is (Pc / Pe) ^{1 / n} . Since the refrigeration apparatuses 10A to 10D include the three-stage compressors 13a, 13b, and 13c, the compression ratio between the compressors is (Pc / Pe) ^1/3 .
Further, in the two-stage isobaric heat release process by the gas coolers 22a and 22b, the discharge gas refrigerant temperature is set to (condensation temperature + α), and the α component is cooled in the isobaric heat release process of each stage, thereby partially compressing Suppresses the discharge temperature from the machine and achieves uniform compression and heat dissipation.

制御装置４２は、再生熱交換器２４の液側出口温度Ｔ_loutが過冷却液出口温度の設定値T_setとなるように、蒸発器２０の入口に直結された第１膨張弁１８の開度を制御する。このときの蒸発器出口の湿り度は状態点（a’〜ａ）の湿り度となるように制御される。再生熱交換器２４のガス側入口状態が状態点（a’〜ａ）であるとき再生熱交換器液側出口温度Ｔ_loutは顕熱限界過冷却点ｈと湿り限界過冷却点ｄの温度範囲内に設定された設定温度T_setと等しくなる。 The controller 42 _opens the opening of the first expansion valve 18 directly connected to the inlet of the evaporator 20 so that the liquid side outlet temperature T _lout of the regenerative heat exchanger 24 becomes the set value T _set of the supercooled liquid outlet temperature. To control. At this time, the wetness of the evaporator outlet is controlled so as to be the wetness of the state points (a ′ to a). When the gas side inlet state of the regenerative heat exchanger 24 is the state point (a ′ to a), the regenerative heat exchanger liquid side outlet temperature T _lout is a temperature range between the sensible heat limit _subcooling point h and the wetness limit _subcooling point d. It becomes equal to the set temperature T _{set set in} .

制御装置４２による例示的な制御方法として、前記第２行程は、電動モータ１４ａ、１４ｂ及び１４ｃの回転数制御と第１膨張弁１８の開度制御とを併用すればよい。
再生熱交換器２４の出口の液冷媒路１２ｃに温度センサ５０が設けられ、蒸発器２０の出口の冷媒路１２に圧力センサ５２及び温度センサ５４が設けられている。
制御装置４２による例示的な制御方法は、まず、凝縮器１６から第１膨張弁１８に向かう液冷媒の再生熱交換器２４の出口における温度を検出する（第１ステップ）。
次に、検出した温度を顕熱限界過冷却点ｈの冷媒温度Ｔｈから温度センサ５４で検出した蒸発器２０における冷媒の蒸発温度までの間の温度に制御する（第２ステップ）。 As an exemplary control method by the control device 42, the second stroke may be performed using both the rotational speed control of the electric motors 14 a, 14 b and 14 c and the opening degree control of the first expansion valve 18.
A temperature sensor 50 is provided in the liquid refrigerant path 12 c at the outlet of the regenerative heat exchanger 24, and a pressure sensor 52 and a temperature sensor 54 are provided in the refrigerant path 12 at the outlet of the evaporator 20.
In the exemplary control method performed by the control device 42, first, the temperature at the outlet of the regenerative heat exchanger 24 of the liquid refrigerant from the condenser 16 toward the first expansion valve 18 is detected (first step).
Next, the detected temperature is controlled to a temperature between the refrigerant temperature Th at the sensible heat limit supercooling point h and the refrigerant evaporation temperature in the evaporator 20 detected by the temperature sensor 54 (second step).

図６及び図８、９において、１段圧縮機１３ａ及び２段圧縮機１３ｂの吐出ガス冷媒（凝縮温度＋α）はガス冷却器２２ａ及び２２ｂにより凝縮温度近傍まで冷却される。３段圧縮機１３ｃの吐出ガス冷媒は凝縮器１６に送られる。
液ガス分離器２８の冷媒温度は膨張弁入口温度Ｔｈから膨張弁出口温度Ｔｋ’まで低下し、第１膨張弁１８を経て蒸発温度Ｔｅとなり、蒸発器２０に供給される。この多段の圧縮・膨張法によりΔＨkfのエンタルピが増加分に対応した全圧縮仕事の変化が起こるが、結果としてＣＯＰが向上する。
図６の近似逆エリクソンサイクルでは、エコノマイザ装置３４のガス冷媒を１段圧縮機１３ａのエコノマイザポートに吸入させることで、第３膨張弁入口温度Ｔｋ’から膨張弁出口温度Ｔl’に低下させ、さらにΔＨlｋのエンタルピが増加し、結果としてＣＯＰを向上できる。 6, 8, and 9, the discharged gas refrigerant (condensation temperature + α) of the first-stage compressor 13 a and the second-stage compressor 13 b is cooled to near the condensation temperature by the gas coolers 22 a and 22 b. The gas refrigerant discharged from the three-stage compressor 13 c is sent to the condenser 16.
The refrigerant temperature of the liquid gas separator 28 decreases from the expansion valve inlet temperature Th to the expansion valve outlet temperature Tk ′, reaches the evaporation temperature Te through the first expansion valve 18, and is supplied to the evaporator 20. This multistage compression / expansion method changes the total compression work corresponding to the increase in the enthalpy of ΔHkf, but as a result, COP is improved.
In the approximate inverse Ericsson cycle of FIG. 6, the gas refrigerant of the economizer device 34 is sucked into the economizer port of the first stage compressor 13a, so that the third expansion valve inlet temperature Tk ′ is lowered to the expansion valve outlet temperature Tl ′. The enthalpy of ΔHlk is increased, and as a result, COP can be improved.

なお、顕熱限界過冷却点ｈの冷媒温度Ｔｈは例えば−１０℃前後になるため、膨張弁出口に付設された中間ガス路３２及びエコノマイザガス路４０は、顕熱限界過冷却点ｈのガス冷媒圧力より低い低段側圧縮機に接続する必要がある。即ち、図６、図８及び図１１等で破線ｍより低圧の領域である。
また、圧縮機の段数を増加するほど吐出温度Ｔｖが低下し、圧縮機の動力を低減できるが、圧縮機段数を増加するほど設備費が増加するので、２段又は３段が適当である。 Since the refrigerant temperature Th at the sensible heat limit supercooling point h is, for example, around −10 ° C., the intermediate gas path 32 and the economizer gas path 40 attached to the outlet of the expansion valve are gas at the sensible heat limit subcooling point h. It is necessary to connect to a lower stage compressor that is lower than the refrigerant pressure. That is, the region is lower than the broken line m in FIGS.
Further, as the number of stages of the compressor is increased, the discharge temperature Tv is decreased, and the power of the compressor can be reduced. However, since the equipment cost increases as the number of compressor stages is increased, two or three stages are appropriate.

冷凍装置１０Ａ〜１０Ｄによれば、過熱蒸気領域で行われる部分行程（等温行程ｂ→ｇ）を、複数段の圧縮機１３ａ、１３ｂ及び１３ｃによる複数段の断熱圧縮行程と、ガス冷却器２２ａ、２２ｂ及び凝縮器１６による複数段の等圧放熱行程とに置き換えることで、圧縮仕事Δを低減でき、ＣＯＰを向上できる。
また、制御装置４２により、蒸発器２０から湿り状態の冷媒を再生熱交換器２４のガス側入口に導入し、該再生熱交換器の液側出口における等温放熱行程の終了点が顕熱限界過冷却点ｈから湿り限界過冷却点ｄまでの領域の過冷却点に位置するように制御することで、冷凍装置１０Ａ〜１０ＤのＣＯＰを最大に維持できる。また、顕熱限界過冷却点ｈと湿り限界過冷却点ｄとの間の広い範囲で制御すればよいので、制御が容易になる。 According to the refrigeration apparatuses 10A to 10D, the partial stroke (isothermal stroke b → g) performed in the superheated steam region is performed by a plurality of stages of adiabatic compression strokes by the plurality of stages of compressors 13a, 13b and 13c, and the gas cooler 22a, By substituting with a multiple-stage isobaric heat release process by 22b and the condenser 16, the compression work Δ can be reduced and the COP can be improved.
Further, the control device 42 introduces the wet refrigerant from the evaporator 20 into the gas side inlet of the regenerative heat exchanger 24, and the end point of the isothermal heat release process at the liquid side outlet of the regenerative heat exchanger is the sensible heat limit exceeded. By controlling so that it may be located in the supercooling point of the area | region from the cooling point h to the wetness limit supercooling point d, COP of freezing apparatus 10A-10D can be maintained to the maximum. Moreover, since control should be performed in a wide range between the sensible heat limit subcooling point h and the wetness limit subcooling point d, the control becomes easy.

また、前記複数段の等圧放熱行程では、（凝縮温度＋α）の吐出ガス冷媒を凝縮温度近傍まで冷却することで、一部の圧縮機での吐出温度の突出を抑制して均質な圧縮と放熱を実現する。
また、過熱蒸気領域における複数段の等圧放熱行程（ｂ’→ｂ）及び（ｂ’→ｇ）によりガス冷媒の温度を低下できるため、冷媒に含まれる潤滑油の劣化及びパッキン材などを構成するエラストマの焼損を防止できる。例えば、凝縮温度４０℃、蒸発温度−４０℃で圧縮機の吸入ガス冷媒温度が４０℃の条件で、比熱比の高い冷媒のＲ７１７（ＮＨ_３）の場合は、単段圧縮では約３８０℃、２段圧縮で約１８０℃、３段圧縮では約１２０℃程度となり、昇温幅αは単段で３４０℃、２段圧縮機で１４０℃、３段圧縮機で８０℃となり、段数が増えることにより冷媒ガス温度の降下が期待できる。
特に、前記複数段の断熱圧縮行程において３段圧縮の場合、圧縮機間の圧縮比を（凝縮圧力Ｐｃ／蒸発圧力Ｐｅ）^１／３としたことで、各圧縮機のガス冷媒の吐出温度をほぼ同等に制御できるため、一部の圧縮機での吐出温度の突出を抑制できる。 In the multiple-stage isobaric heat release process, the discharge gas refrigerant of (condensation temperature + α) is cooled to the vicinity of the condensation temperature, thereby suppressing the discharge temperature from protruding in some compressors and achieving uniform compression. Realize heat dissipation.
In addition, since the temperature of the gas refrigerant can be reduced by a plurality of steps of isobaric heat radiation processes (b ′ → b) and (b ′ → g) in the superheated steam region, the deterioration of the lubricating oil contained in the refrigerant and the packing material are configured. It is possible to prevent the elastomer from burning out. For example, in the case of R717 (NH ₃ ), which is a refrigerant having a high specific heat ratio, under conditions of a condensation temperature of 40 ° C., an evaporation temperature of −40 ° C. and a compressor intake gas refrigerant temperature of 40 ° C., about 380 ° C. in single-stage compression, About 180 ° C for two-stage compression and about 120 ° C for three-stage compression, and the temperature rise width α is 340 ° C for a single stage, 140 ° C for a two-stage compressor, and 80 ° C for a three-stage compressor, increasing the number of stages. Thus, a decrease in refrigerant gas temperature can be expected.
In particular, in the case of three-stage compression in the multiple-stage adiabatic compression stroke, the compression ratio between the compressors is (condensation pressure Pc / evaporation pressure Pe) ^1/3 , so that the discharge temperature of the gas refrigerant of each compressor is Since the control can be performed almost equally, the protrusion of the discharge temperature in some compressors can be suppressed.

また、幾つかの実施形態によれば、中間冷却装置２６又はエコノマイザ装置３４を設けたことで、圧縮機から吐出するガス冷媒の吐出圧力及び吐出温度を低減でき、これによって、圧縮機動力を低減でき、冷凍装置のＣＯＰをさらに向上できる。
また、制御装置４２による前記制御は、複数段の圧縮機１３ａ、１３ｂ及び１３ｃを駆動する電動モータ１４ａ、１４ｂ及び１４ｃの回転数制御と、第１膨張弁１８の開度制御とで行う簡易な制御で容易に行うことができる。
さらに、顕熱限界過冷却点ｈや湿り限界過冷却点ｄの検出は、蒸発器２０の出口や再生熱交換器２４の出口に設けられた温度センサ５０、５４及び圧力センサ５２の検出値から容易に求めることができる。 Further, according to some embodiments, by providing the intermediate cooling device 26 or the economizer device 34, it is possible to reduce the discharge pressure and discharge temperature of the gas refrigerant discharged from the compressor, thereby reducing the compressor power. And COP of the refrigeration apparatus can be further improved.
Further, the control by the control device 42 is simple by performing the rotation speed control of the electric motors 14a, 14b and 14c for driving the compressors 13a, 13b and 13c in a plurality of stages and the opening degree control of the first expansion valve 18. Easy to control.
Further, the detection of the sensible heat limit subcooling point h and the wetness limit subcooling point d is based on the detection values of the temperature sensors 50 and 54 and the pressure sensor 52 provided at the outlet of the evaporator 20 and the outlet of the regenerative heat exchanger 24. It can be easily obtained.

図１３は、他の実施形態として２段圧縮２段膨張を行う冷凍装置１０Ｅを示し、図１４は冷凍装置１０Ｅが構成する近似逆エリクソンサイクルを示している。
前述のように、複数段の圧縮機の段数を２段又は３段とすることで、設備費の増加を抑えつつＣＯＰを向上できる。 FIG. 13 shows a refrigeration apparatus 10E that performs two-stage compression and two-stage expansion as another embodiment, and FIG. 14 shows an approximate inverse Ericsson cycle constituted by the refrigeration apparatus 10E.
As described above, by setting the number of stages of the plurality of compressors to two or three, the COP can be improved while suppressing an increase in equipment costs.

図１５は近似逆エリクソンサイクル及び従来冷凍サイクルのＣＯＰを示す線図である。図１５は中間冷却装置やエコノマイザ装置を付設しない場合を示す。本発明の各実施形態で得られるＣＯＰは従来の冷凍サイクルより向上していることがわかる。
図１５において、冷媒がＲ７１７（ＮＨ_３）で２段圧縮の場合、従来冷凍サイクルよりＣＯＰが下回っているが、圧縮段数を増加するに従い、各圧縮機からのガス冷媒の吐出温度が低下し、圧縮機動力を低減できる。そのため、圧縮損失を改善でき、従来冷凍サイクルよりＣＯＰを向上できる。
また、図１５から、圧縮機の段数が４段以上となってもＣＯＰの増加はあまり見込めないことがわかる。 FIG. 15 is a diagram showing COPs of the approximate inverse Ericsson cycle and the conventional refrigeration cycle. FIG. 15 shows a case where no intermediate cooling device or economizer device is provided. It can be seen that the COP obtained in each embodiment of the present invention is improved over the conventional refrigeration cycle.
In FIG. 15, when the refrigerant is R717 (NH ₃ ) and two-stage compression, the COP is lower than the conventional refrigeration cycle, but as the number of compression stages increases, the discharge temperature of the gas refrigerant from each compressor decreases, Compressor power can be reduced. Therefore, the compression loss can be improved and the COP can be improved compared to the conventional refrigeration cycle.
Further, it can be seen from FIG. 15 that even if the number of compressor stages is four or more, an increase in COP cannot be expected.

図１６は、図５〜図１２に示す前記幾つかの実施形態に係る冷凍装置のＣＯＰ（計算値）を示す図表である。そのうち、図１６比較例No.０５および本発明No.０４は図１５に示す単段圧縮、３段圧縮の近似逆エリクソンサイクルのＣＯＰと同じ値である。
本発明では、中間冷却装置やエコノマイザ装置をさらに備えて顕熱限界過冷却点ｈを下回る冷媒液を、膨張弁を用いて冷媒温度ＴhからＴｋ’またはＴl’に低下させることにより、ΔＨkf（＝ΔＨf−ΔHk）、又はΔＨlf (＝ΔＨf−ΔHl）のエンタルピ差が増加し結果としてＣＯＰが向上することになり、図１６No.１〜No.３の組合せによって更なる性能改善効果が期待できる。 FIG. 16 is a chart showing COPs (calculated values) of the refrigeration apparatus according to some of the embodiments shown in FIGS. Among them, Comparative Example No. 05 in FIG. 16 and No. 04 of the present invention have the same values as the COP of the approximate inverse Ericsson cycle of the single stage compression and the three stage compression shown in FIG.
In the present invention, an intermediate cooling device and an economizer device are further provided to reduce the refrigerant liquid below the sensible heat limit supercooling point h from the refrigerant temperature Th to Tk ′ or Tl ′ using an expansion valve, thereby obtaining ΔHkf (= The difference in enthalpy of ΔHf−ΔHk) or ΔHlf (= ΔHf−ΔHl) increases, and as a result, COP is improved, and a further performance improvement effect can be expected by the combination of No. 1 to No. 3 in FIG.

１０Ａ、１０Ｂ、１０Ｃ、１０Ｄ、１０Ｅ冷凍装置
１２冷媒循環路
１２ａ、１２ｂガス冷媒路
１３ａ、１３ｂ、１３ｃ圧縮機
１４ａ、１４ｂ、１４ｃ電動モータ
１６凝縮器
１８第１膨張弁
２０蒸発器
２２ａ、２２ｂガス冷却器
２４再生熱交換器
２６中間冷却装置
２８、３６液ガス分離器
３０第２膨張弁
３２中間ガス路
３４エコノマイザ装置
３８第３膨張弁
４０エコノマイザガス路
４２制御装置
４４ａ、４４ｂ、４４ｃ回転数センサ
４６、４８流量調整弁
５０、５４温度センサ
５２圧力センサ
Ｔｖ圧縮機吐出温度
ｄ湿り限界過冷却点
ｈ顕熱限界過冷却点
ｐ等圧線
ｔ等温線
ｘ飽和液線
ｙ飽和蒸気線
Δ 圧縮仕事 10A, 10B, 10C, 10D, 10E Refrigerating device 12 Refrigerant circulation path 12a, 12b Gas refrigerant path 13a, 13b, 13c Compressor 14a, 14b, 14c Electric motor 16 Condenser 18 First expansion valve 20 Evaporator 22a, 22b Gas Cooler 24 Regenerative heat exchanger 26 Intermediate cooling device 28, 36 Liquid gas separator 30 Second expansion valve 32 Intermediate gas passage 34 Economizer device 38 Third expansion valve 40 Economizer gas passage 42 Control device 44a, 44b, 44c Speed sensor 46, 48 Flow control valve 50, 54 Temperature sensor 52 Pressure sensor Tv Compressor discharge temperature d Wet limit supercooling point h Sensible heat limit supercooling point p Isobaric line t Isothermal line x Saturated liquid line y Saturated vapor line Δ Compressive work

Claims

冷媒循環路に直列に設けられた複数段の圧縮機、凝縮器、第１膨張手段及び蒸発器と、
前記複数段の圧縮機で各圧縮機の出口と次段の圧縮機の入口とを結ぶ前記冷媒循環路に設けられ、各圧縮機の吐出ガス冷媒を冷却するためのガス冷却器と、
前記蒸発器より前記圧縮機に向かうガス冷媒と前記凝縮器から前記第１膨張手段に向かう液冷媒とを熱交換させるための再生熱交換器と、を備え、
主として前記凝縮器による等温放熱行程と、前記蒸発器による等温吸熱行程と、前記再生熱交換器における熱交換によって行われる液領域における等圧放熱行程及び過熱蒸気領域における等圧吸熱行程とを含む逆エリクソンサイクルを形成すると共に、
前記等温放熱行程のうち過熱蒸気領域で行われる部分行程が、前記複数段の圧縮機による複数段の断熱圧縮行程と、前記ガス冷却器及び前記凝縮器による複数段の等圧放熱行程とに置き換えられ、
前記等圧吸熱行程で湿り状態の冷媒を前記再生熱交換器に導入し、前記等圧吸熱行程で飽和ガス冷媒が前記再生熱交換器に導入されたときの前記等圧放熱行程の終了点となる顕熱限界過冷却点から、前記第１膨張手段による断熱膨張行程の終了点が飽和冷媒液線上に位置するときの前記等圧放熱行程の終了点となる湿り限界過冷却点までの過冷却点に前記等圧放熱行程の終了点が位置するように制御するための制御装置をさらに備えていることを特徴とする冷凍又は空調装置。 A plurality of compressors, condensers, first expansion means and evaporators provided in series in the refrigerant circuit;
A gas cooler for cooling the discharged gas refrigerant of each compressor, provided in the refrigerant circulation path connecting the outlet of each compressor and the inlet of the next stage compressor in the plurality of stages of compressors;
A regenerative heat exchanger for exchanging heat between the gas refrigerant from the evaporator toward the compressor and the liquid refrigerant from the condenser toward the first expansion means,
Inversely, including an isothermal heat release process by the condenser, an isothermal heat absorption process by the evaporator, an isobaric heat release process in the liquid region and an isobaric heat absorption process in the superheated steam region performed by heat exchange in the regeneration heat exchanger. While forming the Ericsson cycle,
The partial process performed in the superheated steam region in the isothermal heat dissipation process is replaced with a plurality of adiabatic compression processes by the multistage compressor and a multistage isobaric heat dissipation process by the gas cooler and the condenser. And
Introducing a wet refrigerant in the regeneration heat exchanger in the isobaric heat absorption process, and an end point of the isobaric heat dissipation process when saturated gas refrigerant is introduced in the regeneration heat exchanger in the isobaric heat absorption process; The subcooling from the sensible heat limit supercooling point to the wetness limit supercooling point which becomes the end point of the isobaric heat release process when the end point of the adiabatic expansion process by the first expansion means is located on the saturated refrigerant liquid line A refrigeration or air conditioner further comprising a control device for controlling the end point of the isobaric heat dissipation process to be located at a point.

前記凝縮器における冷媒の凝縮圧力をＰｃとし、前記蒸発器における冷媒の蒸発圧力をＰｅとし、前記複数段の圧縮機の段数をｎとしたとき、前記複数段の断熱圧縮行程において前記複数段の圧縮機間の圧縮比ｒを（Ｐｃ／Ｐｅ）^１／ｎとし、
前記制御装置は、前記ガス冷却器による前記複数段の等圧放熱行程において、各段の吐出ガス冷媒（凝縮温度＋α）を凝縮温度近傍まで冷却するものであることを特徴とする請求項１に記載の冷凍又は空調装置。 When the condensation pressure of the refrigerant in the condenser is Pc, the evaporation pressure of the refrigerant in the evaporator is Pe, and the number of stages of the plurality of compressors is n, the plurality of stages of adiabatic compression strokes The compression ratio r between the compressors is (Pc / Pe) ^{1 / n} ,
The said control apparatus cools the discharge gas refrigerant | coolant (condensation temperature + (alpha)) of each step | paragraph to the condensing temperature vicinity in the said multiple steps | paragraphs isobaric heat radiation process by the said gas cooler. The refrigeration or air conditioner described.

前記再生熱交換器と前記第１膨張手段との間の前記冷媒循環路に設けられた第１液ガス分離器と、
前記第１液ガス分離器の入口側の前記冷媒循環路に設けられた第２膨張手段と、
前記第１液ガス分離器のガス冷媒を前記複数段の圧縮機のうちの低段側圧縮機の出口と次段の圧縮機の入口に接続された冷媒路に供給する中間ガス路と、を有する中間冷却装置をさらに備えていることを特徴とする請求項１又は２に記載の冷凍又は空調装置。 A first liquid gas separator provided in the refrigerant circuit between the regenerative heat exchanger and the first expansion means;
A second expansion means provided in the refrigerant circuit on the inlet side of the first liquid gas separator;
An intermediate gas passage for supplying gas refrigerant of the first liquid gas separator to a refrigerant passage connected to an outlet of a low-stage compressor of the plurality of stages of compressors and an inlet of a next-stage compressor; The refrigeration or air conditioning apparatus according to claim 1, further comprising an intermediate cooling apparatus having the intermediate cooling apparatus.

前記再生熱交換器と前記第１膨張手段との間の前記冷媒循環路に設けられた第２液ガス分離器と、
前記第２液ガス分離器の入口側の前記冷媒循環路に設けられた第３膨張手段と、
前記第２液ガス分離器のガス冷媒を前記複数段の圧縮機のうちの低段側圧縮機の中間圧領域に供給するエコノマイザガス路と、を有するエコノマイザ装置をさらに備えていることを特徴とする請求項１乃至３の何れか１項に記載の冷凍又は空調装置。 A second liquid gas separator provided in the refrigerant circuit between the regenerative heat exchanger and the first expansion means;
Third expansion means provided in the refrigerant circuit on the inlet side of the second liquid gas separator;
An economizer device having an economizer gas path for supplying a gas refrigerant of the second liquid gas separator to an intermediate pressure region of a low-stage compressor of the plurality of stages of compressors. The refrigeration or air conditioning apparatus according to any one of claims 1 to 3.

前記複数段の圧縮機の段数が２段又は３段であることを特徴とする請求項１乃至４の何れか１項に記載の冷凍又は空調装置。 The refrigeration or air conditioning apparatus according to any one of claims 1 to 4, wherein the number of stages of the plurality of compressors is two or three.

冷媒循環路に直列に設けられた複数段の圧縮機、凝縮器、第１膨張手段及び蒸発器と、
前記複数段の圧縮機で各圧縮機の出口と次段の圧縮機の入口と結ぶ前記冷媒循環路に設けられ、各圧縮機の吐出ガス冷媒を冷却するためのガス冷却器と、
前記蒸発器より前記圧縮機に向かうガス冷媒と前記凝縮器から前記第１膨張手段に向かう液冷媒とを熱交換させるための再生熱交換器と、を備えた冷凍又は空調装置の制御方法において、
主として前記凝縮器による等温放熱行程と、前記蒸発器による等温吸熱行程と、前記再生熱交換器における熱交換によって行われる液領域における等圧放熱行程及び過熱蒸気領域における等圧吸熱行程を含む逆エリクソンサイクルを形成すると共に、
前記等温放熱行程のうち過熱蒸気領域で行われる部分行程が、前記複数段の圧縮機による複数段の断熱圧縮行程と、前記ガス冷却器及び前記凝縮器による複数段の等圧放熱行程とに置き換えた第１行程と、
前記等圧吸熱行程で湿り状態の冷媒を前記再生熱交換器に導入し、前記等圧吸熱行程で飽和ガス冷媒が前記再生熱交換器に導入されたときの前記等温放熱行程の終了点となる顕熱限界過冷却点から、前記第１膨張手段による断熱膨張行程の終了点が飽和冷媒液線上に位置するときの前記等温放熱行程の終了点となる湿り限界過冷却点までの領域の過冷却点に前記等温放熱行程の終了点が位置するように制御する第２行程と、を含むことを特徴とする冷凍又は空調装置の制御方法。 A plurality of compressors, condensers, first expansion means and evaporators provided in series in the refrigerant circuit;
A gas cooler for cooling the discharged gas refrigerant of each compressor, provided in the refrigerant circulation path connecting the outlet of each compressor and the inlet of the next stage compressor in the plurality of stages of compressors;
In a control method of a refrigeration or air conditioner comprising a regenerative heat exchanger for exchanging heat between a gas refrigerant from the evaporator toward the compressor and a liquid refrigerant from the condenser toward the first expansion means,
Inverse Ericsson mainly including an isothermal heat release process in the condenser, an isothermal heat absorption process in the evaporator, an isobaric heat release process in the liquid region and an isobaric heat absorption process in the superheated steam region performed by heat exchange in the regenerative heat exchanger Forming a cycle,
The partial process performed in the superheated steam region in the isothermal heat dissipation process is replaced with a plurality of adiabatic compression processes by the multistage compressor and a multistage isobaric heat dissipation process by the gas cooler and the condenser. The first step,
The wet refrigerant is introduced into the regenerative heat exchanger in the isobaric endothermic process, and is the end point of the isothermal heat releasing process when the saturated gas refrigerant is introduced into the regenerative heat exchanger in the isobaric endothermic process. Supercooling of the region from the sensible heat limit supercooling point to the wetness limit supercooling point which is the end point of the isothermal heat release process when the end point of the adiabatic expansion process by the first expansion means is located on the saturated refrigerant liquid line And a second step of controlling so that the end point of the isothermal heat radiation step is located at a point.

前記第２行程は、少なくとも前記複数段の圧縮機を駆動するモータの回転数制御と、前記第１膨張手段の制御とで行われることを特徴とする請求項６に記載の冷凍又は空調装置の制御方法。 The refrigeration or air conditioning apparatus according to claim 6, wherein the second stroke is performed by at least rotation speed control of a motor that drives the plurality of stages of compressors and control of the first expansion means. Control method.

前記第２行程は、
前記凝縮器から前記第１膨張手段に向かう液冷媒の前記再生熱交換器出口における温度を検出する第１ステップと、
前記第１ステップで検出した温度検出値を、前記顕熱限界過冷却点における冷媒温度から前記蒸発器における冷媒の蒸発温度までの間の温度に制御する第２ステップと、を含むことを特徴とする請求項６又は７に記載の冷凍又は空調装置の制御方法。
The second stroke is
A first step of detecting a temperature at the outlet of the regenerative heat exchanger of liquid refrigerant from the condenser toward the first expansion means;
And a second step of controlling the temperature detection value detected in the first step to a temperature between the refrigerant temperature at the sensible heat limit supercooling point and the evaporation temperature of the refrigerant in the evaporator. The control method of the refrigerating or air-conditioning apparatus of Claim 6 or 7.