JP3985092B2

JP3985092B2 - Air conditioner

Info

Publication number: JP3985092B2
Application number: JP2002066284A
Authority: JP
Inventors: 和幹浦田; 敦彦横関; 眞幸岡部; 福治塚田
Original assignee: Hitachi Ltd
Current assignee: Hitachi Ltd
Priority date: 2002-03-12
Filing date: 2002-03-12
Publication date: 2007-10-03
Anticipated expiration: 2022-03-12
Also published as: JP2003262385A

Description

【０００１】
【発明の属する技術分野】
本発明は空気調和機に係り、特に冷凍サイクルの圧縮機吐出圧力と圧縮機吸入圧力の比である運転圧力比を検出して容量制御するものに好適である。
【０００２】
【従来の技術】
従来、空気調和機において、運転圧力比を求めるため、冷凍サイクルの高圧側及び低圧側に各々圧力センサを設け、その圧力センサの値の比から演算することが知られ、特開平５−１０６０８号公報に記載されている。また、温度センサを冷凍サイクルの低圧側冷媒通路に設け、温度センサ設置部の飽和圧力を温度センサの出力信号より演算し、温度センサ設置部から圧縮機吸入側に至るまでの圧力損失を推定して、圧縮機吸入圧力を演算することが、特開平８−１２１９１６号公報に記載のように知られている。
【０００３】
【発明が解決しようとする課題】
上記従来技術の圧力センサを用いるものにおいては、圧力センサ自体のコストが高く、価格面で不利であり、圧縮機吸入圧力を低圧側冷媒通路に設けた温度センサで推定するものは、室内機と室外機を接続する配管長さが長い場合、圧縮機吸入側の冷媒が過熱状態であるのにも関わらず温度センサ設置部が低い温度となり、吸入圧力の推定値に誤差が生じ、運転圧力比の推定精度が低下する恐れがある。さらに、冷房運転及び暖房運転共に運転圧力比を検出するために温度センサを増やさなければならず、やはり価格面で不利であった。
【０００４】
本発明の目的は、上記従来の技術的課題を解決し、製造コストが安価でかつ運転圧力比を高い精度で検出し、信頼性の高い空気調和機を提供することにある。
【０００５】
【課題を解決するための手段】
上記の目的を達成するために、本発明は、圧縮機、凝縮器、減圧装置、蒸発器とを配管接続して冷凍サイクルを構成し、蒸発器入口温度と凝縮器出口温度とを検出し、前記蒸発器入口温度から前記圧縮機の吸入圧力を、前記凝縮器出口温度から前記圧縮機の吐出圧力をそれぞれ演算して前記圧縮機の吐出圧力と吸入圧力の比である運転圧力比を求めて容量制御する空気調和機において、前記吸入圧力の値は、ガス側接続配管の長さ及び前記圧縮機の運転周波数によって変化する圧力損失を考慮して補正され、この補正された吸入圧力値で前記圧縮機の運転圧力比を求めることを特徴とする。
【０００７】
さらに、上記のものにおいて、凝縮器出口側に開度が可変できる膨脹弁と該膨脹弁の出口に余剰冷媒を貯留する受液器とを設け、膨脹弁を全開して、つまり膨脹弁での圧力損失を小さくするように開度を大きくして運転圧力比を求めることが望ましい。
【０００８】
さらに、上記のものにおいて、吸入圧力は、吸入側の飽和圧力、圧縮機のモータ回転数に比例するものに係数を加えた値、凝縮器と蒸発器を接続する配管長さ、の和となる値のべき乗に関連して求めることが望ましい。
【０００９】
さらに、上記のものにおいて、求めた運転圧力比が所定の範囲となるように圧縮機の運転周波数を制御することが望ましい。
【００１０】
【発明の実施の形態】
以下、本発明の一実施の形態を図１ないし図４を参照して説明する。
図１は、冷凍サイクル構成を示し、室外機１００と室内機２００に大別され、室外機１００と室内機２００は液側接続配管５１及びガス側接続配管５０で接続されている。室外機１００は圧縮機１、四方弁２、室外熱交換器４、室外ファン５より構成され、室外熱交換器４には凝縮器として作用した場合に出口側、蒸発器として作用した場合に入口側となる位置に温度センサ２１が設けられている。また、室内機２００は室内熱交換器３、室内膨張弁７、室内ファン６より構成され、室内熱交換器３には、室外熱交換器４と同様に凝縮器として作用した場合に出口側、蒸発器として作用した場合に入口側となる位置に温度センサ２２が設けられている。室内熱交換器３及び室外熱交換器４に設けられている温度センサの信号は、マイクロコンピュータ２０内に入力されるように配線されている。
【００１１】
次に、冷凍サイクルの運転動作について説明する。図１において、実線矢印は冷房運転での冷媒の流れ方向を示し、破線矢印は暖房運転時での冷媒の流れ方向を示す。
冷房運転の場合、圧縮機１で圧縮された高温高圧のガス冷媒は、四方弁２を通り室外熱交換器４に流入し、室外ファン５により室外熱交換器４に送られる空気と熱交換して凝縮液化して室外熱交換器４から流出する。凝縮液化した冷媒は、液側接続配管５１を通り室内膨張弁７で減圧され気液二相状態の冷媒となり、室内熱交換器３に流入し、室内ファン６により室内熱交換器３に送られる空気と熱交換して蒸発ガス化して室内熱交換器３から流出する。蒸発ガス化した冷媒は、ガス接続配管５０、四方弁２を通り圧縮機１に戻り、再び圧縮機１で圧縮されることで冷凍サイクルが形成される。
【００１２】
図２は、冷凍サイクルモリエル線図を表わし、点ａは圧縮機１から吐出された状態を示しており、その圧力である吐出圧力はＰｄとなる。点ｂは室外熱交換器４出口すなわち凝縮器出口の状態であり、点ｃは室内膨張弁７で減圧された室内熱交換器３入口すなわち蒸発器入口の状態を示す。点ｄは圧縮機１に吸入される状態を示しており、その圧力である吸入圧力はＰｓとなる。冷房運転時の運転圧力比εは、吐出圧力Ｐｄと吸入圧力Ｐｓの比となる。
【００１３】
圧縮機の運転可能範囲として運転圧力比εは、λ１（例えば、１．８）＜ε＜λ２（例えば、８．０）が望ましく、運転圧力比εがλ１以下、あるいはλ２以上となると圧縮機の運転寿命が極端に減少し、早期に故障を起こす恐れがある。このため、運転圧力比εがλ１とλ２の間になるように、圧縮機周波数、室内外ファンや室内外の膨張弁開度を制御して、圧縮機の信頼性を確保する。つまり、運転圧力比ε＜λ１となった場合は、圧縮機周波数の下限値を通常制御時よりも高くし、冷房運転であれば室外ファンの風量を減少させ、暖房運転であれば室内ファンの風量を増加させる。また、運転圧力比ε＞λ２となった場合は、圧縮機周波数の上限値を通常制御時よりも低くし、冷房運転であれば室外ファンの風量を増加させ、暖房運転であれば室外ファンの風量を減少させたり、室内膨張弁の開度を開けて凝縮器として作用している熱交換器内の冷媒をレシーバ等に回収させたりする。
【００１４】
次に、冷房運転時の運転圧力比の推定方法について説明する。運転圧力比の推定は、室内熱交換器３及び室外熱交換器４に設けた温度センサ２１，２２及び圧縮機１のモータ回転数の信号を用いてマイクロコンピュータ２０で行う。
運転圧力比εは、室外熱交換器４に設けた温度センサ２１の検出値Ｔｃから推定した吐出圧力Ｐｄと、室内熱交換器３に設けた温度センサ２２の検出値Ｔｅから推定した吸入圧力Ｐｓとの比となる。
【００１５】
【数１】

【００１６】
推定吐出圧力Ｐｄは自然対数の底ｅの｛α２×（Ｔｃ＋α３）｝乗に比例するとし、推定吸入圧力Ｐｓは、底ｅの｛β２×（Ｔｅ＋Ｆ１＋Ｌ１）｝乗に比例するとする。
【００１７】
【数２】

【００１８】
【数３】

【００１９】
Ｔｃは、吐出側の飽和圧力、Ｔｅは、吸入側の飽和圧力であり、α１〜α３、β１、β２、は空気調和機に封入する冷媒の種類に応じて変わる係数である。Ｆ１はモータ回転数Ｈｚによる補正値であり、Ｌ１は、凝縮器と蒸発器又は室外機と室内機とを接続する接続配管長さによる補正係数である。Ｆ１は数４に示すように圧縮機１のモータ回転数の検出値Ｈｚに比例するものにγ２を加えたものとする。
【００２０】
【数４】

【００２１】
Ｌ１は、室内機２００と室外機１００を接続する配管長さによって変わり、マイクロコンピュータ２０に予め入力されているか、ディップスイッチや通信等による外部信号から入力される。
【００２２】
図２に示すモリエル線図で説明すると、室外熱交換器４に設けられている温度センサ２１により飽和圧力Ｔｃが求められ、Ｔｃを数２により求めて推定吐出圧力Ｐｄが求められる。推定Ｐｄは、実際の吐出圧力であるＰｄと比べて四方弁２や室外熱交換器４の圧力損失及び凝縮器の冷媒過冷却度により若干低くなるが、冷媒の状態が高温高圧で密度が大きく圧力損失が小さいこと、及び温度センサまでに至る距離が変化しないことから、数１の係数を調整することで推定Ｐｄと実際のＰｄを同値とすることができる。
【００２３】
一方、室内熱交換器３に設けられている温度センサ２２により吸入側の飽和圧力Ｔｅが求められる。飽和圧力Ｔｅは、室内熱交換器３やガス接続配管５０等での圧力損失があるため吸入圧力Ｐｓより高くなる。この圧力損失は、ガス側接続配管５０の長さ及び冷凍サイクルを流れる冷媒循環量で変化するため、数４に示すように補正する。つまり、冷凍サイクルを流れる冷媒循環量は圧縮機１のモータ回転数Ｈｚに比例することから、推定吸入圧力Ｐｓは接続配管長さによる補正係数Ｌ１と圧縮機１のモータ回転数Ｈｚによる補正値Ｆ１により求める。
【００２４】
暖房運転の場合、圧縮機１で圧縮された高温高圧のガス冷媒は、四方弁２、ガス側接続配管５０を通り室内熱交換器３に流入し、室内ファン６により室内熱交換器３に送られる空気と熱交換して凝縮液化して室内熱交換器３から流出する。凝縮液化した冷媒は、室内膨張弁７で減圧され気液二相状態の冷媒となり、液側接続配管５１を通り室外熱交換器４に流入し、室外ファン５により室外熱交換器４に送られる空気と熱交換して蒸発ガス化して室外熱交換器４から流出する。蒸発ガス化した冷媒は、四方弁２を通り圧縮機１に戻り、再び圧縮機１で圧縮されることで冷凍サイクルが形成される。
【００２５】
暖房運転の冷凍サイクルをモリエル線図に示すと、冷房運転と同様に図２となり、暖房運転時の運転圧力比は、冷房運転時と同様に求めることができる。また、暖房運転時の運転圧力比の推定は、冷房運転時と同様に数１ないし４を用いて行う。ただし、数式中のＴｃは室内熱交換器３に設けた温度センサ２２の検出値、Ｔｅは室外熱交換器４に設けた温度センサ２１の検出値とする。
【００２６】
以上に示すように、凝縮器出口温度及び蒸発器入口温度を検出することにより空気調和機の運転圧力比の推定が可能であり、高価な圧力センサを使用せずに運転圧力比を検出でき、信頼性を向上できる。また、吐出圧力及び吸入圧力を温度センサ（サーミスタ）で推定するため、吐出圧力推定値から吐出ガスの飽和温度を推定することが可能であるため、吐出温度を検出することで吐出ガス過熱度を推定することが可能であり、吐出ガス過熱度を膨張弁で制御することができる。さらに、吸入圧力の推定値を用いて、圧縮機の真空運転を防止するように冷凍サイクルの制御機器を制御できるため、圧縮機の信頼性をより一層向上することができる。
【００２７】
次に、本発明の他の実施形態について図３ないし４を参照して説明する。
図３は、凝縮器の冷媒過冷却度制御装置を具備した空気調和機の冷凍サイクルを示し、冷凍サイクルは、室内膨張弁７と室外熱交換器４の間に受液器９と室外膨張弁８を配置した構成となっている。
【００２８】
受液器９は、余剰冷媒が貯留され、冷房運転時は室外膨張弁８を全開で、つまり膨脹弁での圧力損失を小さくするように開度を大きくして使用し、室内膨張弁７で冷凍サイクルの温度を制御し、暖房運転時は室内膨張弁７を全開で使用し室外膨張弁８で冷凍サイクルの温度を制御する。冷凍サイクルの運転をモリエル線図で示すと図４のようになる。すなわち、凝縮器として作用している熱交換器の出口に設けられている膨張弁を全開で使用するため、凝縮器冷媒過冷却度は０となり、凝縮器出口の温度センサで検知したＴｃの飽和圧力は吐出圧力Ｐｄと同値となる。よって、受液器９を設け、凝縮器として作用している熱交換器の出口に設けられている膨張弁を全開にして凝縮器冷媒過冷却度を制御することで、吐出圧力Ｐｄの推定値の精度を向上することができ、空気調和機の運転圧力比をより精度良くして、信頼性を向上できる。
【００２９】
【発明の効果】
以上説明したように、本発明によれば、高価な圧力センサを使用せずに運転圧力比を検出し、空気調和機の製造コストを大幅に低減できると共に、信頼性の向上を図ることができる。
【図面の簡単な説明】
【図１】本発明の一実施形態を示す空気調和機の冷凍サイクル。
【図２】図１の冷凍サイクルの運転状態を示したモリエル線図。
【図３】他の実施形態を示す空気調和機の冷凍サイクル。
【図４】図２の冷凍サイクルの運転状態を示したモリエル線図。
【符号の説明】
１…圧縮機、３…室内熱交換器、４…室外熱交換器、７…室内膨張弁、８…室外膨張弁、９…受液器、２０…マイクロコンピュータ、２１，２２…温度センサ、５０…ガス側接続配管、５１…液側接続配管、１００…室外機、２００…室内機。[0001]
BACKGROUND OF THE INVENTION
The present invention relates to an air conditioner, and is particularly suitable for detecting the operating pressure ratio, which is the ratio of the compressor discharge pressure and the compressor suction pressure in the refrigeration cycle, and controlling the capacity.
[0002]
[Prior art]
Conventionally, in an air conditioner, in order to obtain an operating pressure ratio, it is known to provide a pressure sensor on each of a high pressure side and a low pressure side of a refrigeration cycle and calculate from the ratio of the values of the pressure sensors. It is described in the publication. In addition, a temperature sensor is provided in the low-pressure side refrigerant passage of the refrigeration cycle, the saturation pressure of the temperature sensor installation part is calculated from the output signal of the temperature sensor, and the pressure loss from the temperature sensor installation part to the compressor suction side is estimated. Thus, it is known to calculate the compressor suction pressure as described in JP-A-8-121916.
[0003]
[Problems to be solved by the invention]
Among those using the above prior art pressure sensor, the cost of the pressure sensor itself is high and disadvantageous in terms of price, and the one that estimates the compressor suction pressure with the temperature sensor provided in the low-pressure side refrigerant passage is an indoor unit. If the length of the pipe connecting the outdoor unit is long, the temperature sensor installation part will be at a low temperature even though the refrigerant on the compressor suction side is overheated, causing an error in the estimated value of the suction pressure, and the operating pressure ratio There is a risk that the estimation accuracy of will decrease. Furthermore, the temperature sensor has to be increased in order to detect the operating pressure ratio in both the cooling operation and the heating operation, which is also disadvantageous in terms of price.
[0004]
An object of the present invention is to solve the above-described conventional technical problems, to provide an air conditioner that has a low manufacturing cost, detects an operating pressure ratio with high accuracy, and has high reliability.
[0005]
[Means for Solving the Problems]
In order to achieve the above-mentioned object, the present invention configures a refrigeration cycle by connecting a compressor, a condenser, a decompression device, and an evaporator, and detects an evaporator inlet temperature and a condenser outlet temperature. The compressor suction pressure is calculated from the evaporator inlet temperature, and the compressor discharge pressure is calculated from the condenser outlet temperature to obtain an operating pressure ratio that is the ratio of the compressor discharge pressure to the suction pressure. In the air conditioner for capacity control, the value of the suction pressure is corrected in consideration of the pressure loss that varies depending on the length of the gas side connection pipe and the operating frequency of the compressor, and the corrected suction pressure value The operation pressure ratio of the compressor is obtained.
[0007]
Further, in the above, an expansion valve whose opening degree can be varied on the outlet side of the condenser and a receiver for storing excess refrigerant at the outlet of the expansion valve are provided, and the expansion valve is fully opened, that is, the expansion valve It is desirable to obtain the operating pressure ratio by increasing the opening so as to reduce the pressure loss.
[0008]
Further, in the above, the suction pressure is the sum of the saturation pressure on the suction side, the value proportional to the motor speed of the compressor plus a coefficient, and the length of the pipe connecting the condenser and the evaporator. It is desirable to find it in relation to the power of the value.
[0009]
Furthermore, in the above, it is desirable to control the operating frequency of the compressor so that the obtained operating pressure ratio falls within a predetermined range.
[0010]
DETAILED DESCRIPTION OF THE INVENTION
An embodiment of the present invention will be described below with reference to FIGS.
FIG. 1 shows a refrigeration cycle configuration, which is roughly divided into an outdoor unit 100 and an indoor unit 200, and the outdoor unit 100 and the indoor unit 200 are connected by a liquid side connection pipe 51 and a gas side connection pipe 50. The outdoor unit 100 includes a compressor 1, a four-way valve 2, an outdoor heat exchanger 4, and an outdoor fan 5. The outdoor heat exchanger 4 has an outlet side when acting as a condenser, and an inlet when acting as an evaporator. A temperature sensor 21 is provided at a position on the side. The indoor unit 200 includes an indoor heat exchanger 3, an indoor expansion valve 7, and an indoor fan 6, and the indoor heat exchanger 3 has an outlet side when acting as a condenser, like the outdoor heat exchanger 4, A temperature sensor 22 is provided at a position on the inlet side when acting as an evaporator. Signals from temperature sensors provided in the indoor heat exchanger 3 and the outdoor heat exchanger 4 are wired so as to be input into the microcomputer 20.
[0011]
Next, the operation of the refrigeration cycle will be described. In FIG. 1, a solid line arrow indicates the flow direction of the refrigerant in the cooling operation, and a broken line arrow indicates the flow direction of the refrigerant in the heating operation.
In the cooling operation, the high-temperature and high-pressure gas refrigerant compressed by the compressor 1 flows into the outdoor heat exchanger 4 through the four-way valve 2 and exchanges heat with air sent to the outdoor heat exchanger 4 by the outdoor fan 5. Then, it is liquefied and flows out of the outdoor heat exchanger 4. The condensed and liquefied refrigerant passes through the liquid side connection pipe 51 and is decompressed by the indoor expansion valve 7 to become a gas-liquid two-phase refrigerant, flows into the indoor heat exchanger 3, and is sent to the indoor heat exchanger 3 by the indoor fan 6. It exchanges heat with air to evaporate gas and flows out from the indoor heat exchanger 3. The evaporated gas refrigerant returns to the compressor 1 through the gas connection pipe 50 and the four-way valve 2 and is compressed again by the compressor 1 to form a refrigeration cycle.
[0012]
FIG. 2 shows a refrigeration cycle Mollier diagram. A point a indicates a state of being discharged from the compressor 1, and a discharge pressure which is the pressure is Pd. Point b indicates the state of the outlet of the outdoor heat exchanger 4, that is, the outlet of the condenser, and point c indicates the state of the inlet of the indoor heat exchanger 3, that is, the inlet of the evaporator, which is decompressed by the indoor expansion valve 7. A point d indicates a state where the air is sucked into the compressor 1, and the suction pressure that is the pressure is Ps. The operating pressure ratio ε during cooling operation is the ratio between the discharge pressure Pd and the suction pressure Ps.
[0013]
The operating pressure ratio ε is preferably λ1 (for example, 1.8) <ε <λ2 (for example, 8.0) as an operable range of the compressor. When the operating pressure ratio ε is λ1 or less, or λ2 or more, the compressor The operating life of the battery may be drastically reduced, leading to failure at an early stage. For this reason, the reliability of the compressor is ensured by controlling the compressor frequency, the indoor / outdoor fan and the indoor / outdoor expansion valve opening so that the operating pressure ratio ε is between λ1 and λ2. In other words, when the operating pressure ratio ε <λ1, the lower limit value of the compressor frequency is set higher than that during normal control, the air volume of the outdoor fan is reduced during cooling operation, and the indoor fan Increase air flow. Also, when the operating pressure ratio ε> λ2, the upper limit value of the compressor frequency is set lower than that during normal control, the air volume of the outdoor fan is increased during cooling operation, and the outdoor fan The air volume is reduced, or the opening of the indoor expansion valve is opened, and the refrigerant in the heat exchanger acting as a condenser is collected by a receiver or the like.
[0014]
Next, a method for estimating the operating pressure ratio during cooling operation will be described. The operation pressure ratio is estimated by the microcomputer 20 using the

temperature sensors

21 and 22 provided in the indoor heat exchanger 3 and the outdoor heat exchanger 4 and the motor rotation speed signal of the compressor 1.
The operating pressure ratio ε is the discharge pressure Pd estimated from the detection value Tc of the temperature sensor 21 provided in the outdoor heat exchanger 4 and the suction pressure Ps estimated from the detection value Te of the temperature sensor 22 provided in the indoor heat exchanger 3. And the ratio.
[0015]
[Expression 1]

[0016]
Assume that the estimated discharge pressure Pd is proportional to the power of {α2 × (Tc + α3)} of the natural logarithm e, and the estimated suction pressure Ps is proportional to the power of {β2 × (Te + F1 + L1)} of the base e.
[0017]
[Expression 2]

[0018]
[Equation 3]

[0019]
Tc is the saturation pressure on the discharge side, Te is the saturation pressure on the suction side, and α1 to α3, β1, and β2 are coefficients that vary depending on the type of refrigerant sealed in the air conditioner. F1 is a correction value based on the motor rotation speed Hz, and L1 is a correction coefficient based on the length of a connection pipe connecting the condenser and the evaporator or the outdoor unit and the indoor unit. As shown in Formula 4, F1 is assumed to be obtained by adding γ2 to a value proportional to the detected value Hz of the motor rotation number of the compressor 1.
[0020]
[Expression 4]

[0021]
L1 varies depending on the length of the pipe connecting the indoor unit 200 and the outdoor unit 100, and is input to the microcomputer 20 in advance or input from an external signal such as a dip switch or communication.
[0022]
Referring to the Mollier diagram shown in FIG. 2, the saturation pressure Tc is obtained by the temperature sensor 21 provided in the outdoor heat exchanger 4, and the estimated discharge pressure Pd is obtained by obtaining Tc by Equation 2. The estimated Pd is slightly lower than Pd, which is the actual discharge pressure, due to the pressure loss of the four-way valve 2 and the outdoor heat exchanger 4 and the degree of refrigerant supercooling of the condenser. Since the pressure loss is small and the distance to the temperature sensor does not change, the estimated Pd and the actual Pd can be made equal by adjusting the coefficient of Equation 1.
[0023]
On the other hand, the saturation pressure Te on the suction side is obtained by the temperature sensor 22 provided in the indoor heat exchanger 3. The saturation pressure Te is higher than the suction pressure Ps because of pressure loss in the indoor heat exchanger 3, the gas connection pipe 50, and the like. Since this pressure loss varies depending on the length of the gas side connection pipe 50 and the amount of refrigerant circulating through the refrigeration cycle, the pressure loss is corrected as shown in Equation 4. That is, since the amount of refrigerant circulating through the refrigeration cycle is proportional to the motor rotation speed Hz of the compressor 1, the estimated suction pressure Ps is a correction value L1 based on the correction coefficient L1 due to the connection pipe length and the motor rotation speed Hz of the compressor 1. Ask for.
[0024]
In the case of heating operation, the high-temperature and high-pressure gas refrigerant compressed by the compressor 1 flows into the indoor heat exchanger 3 through the four-way valve 2 and the gas side connection pipe 50, and is sent to the indoor heat exchanger 3 by the indoor fan 6. Heat exchange with the air to be condensed and liquefied to flow out of the indoor heat exchanger 3. The condensed and liquefied refrigerant is decompressed by the indoor expansion valve 7 to become a gas-liquid two-phase refrigerant, flows through the liquid side connection pipe 51 and flows into the outdoor heat exchanger 4, and is sent to the outdoor heat exchanger 4 by the outdoor fan 5. It exchanges heat with air to evaporate gas and flows out of the outdoor heat exchanger 4. The evaporated gas refrigerant returns to the compressor 1 through the four-way valve 2 and is compressed again by the compressor 1 to form a refrigeration cycle.
[0025]
When the refrigeration cycle of the heating operation is shown in the Mollier diagram, it becomes FIG. 2 like the cooling operation, and the operation pressure ratio during the heating operation can be obtained in the same manner as during the cooling operation. Further, the estimation of the operating pressure ratio during the heating operation is performed using Equations 1 to 4 as in the cooling operation. In the equation, Tc is a detection value of the temperature sensor 22 provided in the indoor heat exchanger 3 and Te is a detection value of the temperature sensor 21 provided in the outdoor heat exchanger 4.
[0026]
As shown above, the operating pressure ratio of the air conditioner can be estimated by detecting the condenser outlet temperature and the evaporator inlet temperature, and the operating pressure ratio can be detected without using an expensive pressure sensor, Reliability can be improved. Moreover, since the discharge pressure and the suction pressure are estimated by a temperature sensor (thermistor), it is possible to estimate the saturation temperature of the discharge gas from the estimated discharge pressure value. Therefore, by detecting the discharge temperature, the discharge gas superheat degree can be determined. It is possible to estimate, and the discharge gas superheat degree can be controlled by the expansion valve. Furthermore, since the estimated value of the suction pressure can be used to control the control device of the refrigeration cycle so as to prevent the vacuum operation of the compressor, the reliability of the compressor can be further improved.
[0027]
Next, another embodiment of the present invention will be described with reference to FIGS.
FIG. 3 shows a refrigeration cycle of an air conditioner equipped with a refrigerant supercooling degree control device for a condenser. The refrigeration cycle includes a receiver 9 and an outdoor expansion valve between the indoor expansion valve 7 and the outdoor heat exchanger 4. 8 is arranged.
[0028]
The receiver 9 stores excess refrigerant, and is used with the outdoor expansion valve 8 fully opened during cooling operation, that is, with the opening degree increased so as to reduce the pressure loss at the expansion valve. The temperature of the refrigeration cycle is controlled. During the heating operation, the indoor expansion valve 7 is fully opened and the outdoor expansion valve 8 controls the temperature of the refrigeration cycle. FIG. 4 shows the operation of the refrigeration cycle as a Mollier diagram. That is, since the expansion valve provided at the outlet of the heat exchanger acting as a condenser is used fully open, the condenser refrigerant supercooling degree becomes 0, and the saturation of Tc detected by the temperature sensor at the condenser outlet. The pressure has the same value as the discharge pressure Pd. Therefore, the estimated value of the discharge pressure Pd is provided by providing the liquid receiver 9 and fully opening the expansion valve provided at the outlet of the heat exchanger acting as a condenser to control the degree of subcooling of the condenser refrigerant. The accuracy of the air conditioner can be improved, the operating pressure ratio of the air conditioner can be improved more accurately, and the reliability can be improved.
[0029]
【The invention's effect】
As described above, according to the present invention, it is possible to detect the operating pressure ratio without using an expensive pressure sensor, greatly reduce the manufacturing cost of the air conditioner, and improve the reliability. .
[Brief description of the drawings]
FIG. 1 is a refrigeration cycle of an air conditioner showing an embodiment of the present invention.
FIG. 2 is a Mollier diagram showing the operating state of the refrigeration cycle of FIG.
FIG. 3 is a refrigeration cycle of an air conditioner showing another embodiment.
4 is a Mollier diagram showing the operating state of the refrigeration cycle of FIG. 2;
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 1 ... Compressor, 3 ... Indoor heat exchanger, 4 ... Outdoor heat exchanger, 7 ... Indoor expansion valve, 8 ... Outdoor expansion valve, 9 ... Liquid receiver, 20 ... Microcomputer, 21,22 ... Temperature sensor, 50 ... gas side connection piping, 51 ... liquid side connection piping, 100 ... outdoor unit, 200 ... indoor unit.

Claims

圧縮機、凝縮器、減圧装置、蒸発器とを配管接続して冷凍サイクルを構成し、蒸発器入口温度と凝縮器出口温度とを検出し、前記蒸発器入口温度から前記圧縮機の吸入圧力を、前記凝縮器出口温度から前記圧縮機の吐出圧力をそれぞれ演算して前記圧縮機の吐出圧力と吸入圧力の比である運転圧力比を求めて容量制御する空気調和機において、
前記吸入圧力の値は、ガス側接続配管の長さ及び前記圧縮機の運転周波数によって変化する圧力損失を考慮して補正され、
この補正された吸入圧力値で前記圧縮機の運転圧力比を求めることを特徴とする空気調和機。A compressor, a condenser, a decompression device, and an evaporator are connected to each other to form a refrigeration cycle , and an evaporator inlet temperature and a condenser outlet temperature are detected, and an intake pressure of the compressor is determined from the evaporator inlet temperature. In the air conditioner for controlling the capacity by calculating the discharge pressure of the compressor from the outlet temperature of the condenser to obtain the operating pressure ratio which is the ratio of the discharge pressure and the suction pressure of the compressor ,
The value of the suction pressure is corrected in consideration of the pressure loss that varies depending on the length of the gas side connection pipe and the operating frequency of the compressor,
An air conditioner characterized in that an operating pressure ratio of the compressor is obtained from the corrected suction pressure value .

請求項１に記載のものにおいて、前記凝縮器出口側に開度が可変できる膨脹弁と該膨脹弁の出口に余剰冷媒を貯留する受液器とを設け、前記膨脹弁を全開して前記運転圧力比を求めることを特徴とする空気調和機。2. The operation according to claim 1, further comprising: an expansion valve whose opening degree can be varied on the outlet side of the condenser; and a liquid receiver that stores excess refrigerant at an outlet of the expansion valve, and the expansion valve is fully opened to perform the operation. An air conditioner characterized by obtaining a pressure ratio .

請求項１に記載のものにおいて、求めた運転圧力比が所定の範囲となるように前記圧縮機の運転周波数を制御することを特徴とする空気調和機。2. The air conditioner according to claim 1, wherein the operating frequency of the compressor is controlled so that the obtained operating pressure ratio falls within a predetermined range .