JPH0120692B2 - - Google Patents
Info
- Publication number
- JPH0120692B2 JPH0120692B2 JP10765783A JP10765783A JPH0120692B2 JP H0120692 B2 JPH0120692 B2 JP H0120692B2 JP 10765783 A JP10765783 A JP 10765783A JP 10765783 A JP10765783 A JP 10765783A JP H0120692 B2 JPH0120692 B2 JP H0120692B2
- Authority
- JP
- Japan
- Prior art keywords
- flow rate
- pressure
- tube
- pressure reducer
- differential pressure
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Expired
Links
- 239000003638 chemical reducing agent Substances 0.000 claims description 16
- 238000002347 injection Methods 0.000 claims description 8
- 239000007924 injection Substances 0.000 claims description 8
- 239000007788 liquid Substances 0.000 claims description 7
- 238000005057 refrigeration Methods 0.000 claims description 4
- 230000007423 decrease Effects 0.000 description 10
- 238000001816 cooling Methods 0.000 description 8
- 230000000694 effects Effects 0.000 description 2
- 239000003507 refrigerant Substances 0.000 description 2
- 238000010586 diagram Methods 0.000 description 1
Classifications
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
- F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
- F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
- F25B2400/13—Economisers
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
- F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
- F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
- F25B2400/23—Separators
Description
【発明の詳細な説明】
〔発明の利用分野〕
本発明はパツケージ形空調機、ルームエアコン
などのガスインジエクシヨン回路を備えた冷凍サ
イクルに関するものである。DETAILED DESCRIPTION OF THE INVENTION [Field of Application of the Invention] The present invention relates to a refrigeration cycle equipped with a gas injection circuit for a package air conditioner, a room air conditioner, or the like.
第１図に示されるような第２減圧器４にキヤピ
ラリチユーブを用いたガスインジエクシヨンサイ
クルでは、凝縮温度が低下（例えば、冷房運転で
外気温度が低下）して第２減圧器入口出口の圧力
差が小さくなつた場合、キヤピラリチユーブでの
流量特性が変化して、サイクル各部の圧力が周期
的に変動する状況が生じたり、極度に流量が減少
して能力不足の状態となる。
In a gas injection cycle using a capillary tube as the second pressure reducer 4 as shown in FIG. If the pressure difference becomes small, the flow characteristics in the capillary tube change, causing a situation in which the pressure in each part of the cycle fluctuates periodically, or the flow rate decreases extremely, resulting in a state of insufficient capacity.
これは第２減圧器であるキヤピラリチユーブの
流量特性によるものである。第２図にキヤピラリ
チユーブの流量特性を示す。図は入口圧力を一定
として出口圧力を変化させた場合の入口出口の差
圧ΔPと流量Ｇの関係であり、入口圧力がパラメ
ータとなつている。入口圧力P1の場合を例にと
ると、差圧ΔPが小さい領域では、差圧ΔPの増加
とともに流量Ｇは増加するが、ある差圧ΔP1以上
では、出口部でチヨーク現象を起こすために差圧
ΔPに対して流量Ｇはほぼ一定（最大流量G1）と
なる。また入口圧力が低くなると流量は小さくな
るが、この傾向は同様である。第３図は冷凍サイ
クルにおいて、凝縮器側の負荷が変動、例えば冷
房運転で外気温度が変化した場合の第２減圧器
（キヤピラリチユーブ）での作動点を示す。標準
の設計点においては、差圧ΔP2に対して最大流量
G2が得られるようにキヤピラリチユーブの寸法
諸元（管直径及び長さ）が決定される。凝縮温度
が低下すると、キヤピラリチユーブ入口圧力は
P1，P0と低下し、キヤピラリチユーブ入口出口
の差圧ΔPもΔP1，ΔP2と低下する。入口圧力が
P1の場合には差圧ΔP1で最大流量G1が得られる
点でバランスするが、入口圧力がP0まで低下し
た場合には、差圧はΔP0となり、P0に対する最大
流量が得られず、これよりも少ない流量G0とな
る。この場合には、差圧ΔPに対して流量Ｇが変
化する特性となるため、空気温度の少しの変動な
どにより、キヤピラリチユーブ入口出口の差圧が
変化して流量が変動するため、冷凍サイクルの運
転点は不安定となり易い。また、ΔP0での流量G0
は、最大流量よりも小さいため、蒸発器出口冷媒
の過熱度が大きい運転となる。このような運転で
は、運転効率が悪く、また圧力変動が大きい場合
には、間欠的に液戻りが生じることから圧縮機の
信頼性低下につながる。 This is due to the flow characteristics of the capillary tube, which is the second pressure reducer. Figure 2 shows the flow characteristics of the capillary tube. The figure shows the relationship between the pressure difference ΔP at the inlet and outlet and the flow rate G when the inlet pressure is kept constant and the outlet pressure is varied, and the inlet pressure is the parameter. Taking the case of an inlet pressure P 1 as an example, in a region where the differential pressure ΔP is small, the flow rate G increases as the differential pressure ΔP increases, but at a certain differential pressure ΔP 1 or more, the flow rate G increases due to the occurrence of the chyoke phenomenon at the outlet. The flow rate G is approximately constant (maximum flow rate G 1 ) with respect to the differential pressure ΔP. Furthermore, as the inlet pressure decreases, the flow rate decreases, but this trend remains the same. FIG. 3 shows operating points in the second pressure reducer (capillary tube) when the load on the condenser side changes, for example, when the outside air temperature changes during cooling operation in the refrigeration cycle. At standard design point, maximum flow rate for differential pressure ∆P 2
The dimensions of the capillary tube (tube diameter and length) are determined so that G 2 is obtained. As the condensing temperature decreases, the capillary tube inlet pressure becomes
The pressure difference ΔP at the entrance and exit of the capillary tube also decreases to ΔP 1 and ΔP 2 . The inlet pressure
In the case of P 1 , there is a balance in that the maximum flow rate G 1 is obtained with a differential pressure ΔP 1 , but if the inlet pressure drops to P 0 , the differential pressure becomes ΔP 0 and the maximum flow rate with respect to P 0 is obtained. Therefore, the flow rate G 0 will be lower than this. In this case, since the flow rate G changes with respect to the differential pressure ΔP, the differential pressure at the inlet and outlet of the capillary tube changes due to slight fluctuations in air temperature, causing the flow rate to fluctuate. The operating point of is likely to become unstable. Also, the flow rate G 0 at ΔP 0
Since this is smaller than the maximum flow rate, the refrigerant at the evaporator outlet is operated with a high degree of superheating. In such an operation, the operating efficiency is poor, and when pressure fluctuations are large, liquid returns occur intermittently, leading to a decrease in the reliability of the compressor.
本発明は、凝縮温度が低く、第２減圧器入口出
口での圧力差が小さい場合でも、差圧に対して流
量が変動しない最大流量が得られるような流量特
性を有する第２減圧器を提供することを目的とす
るものである。
The present invention provides a second pressure reducer having flow characteristics such that even when the condensing temperature is low and the pressure difference at the inlet and outlet of the second pressure reducer is small, a maximum flow rate is obtained in which the flow rate does not fluctuate with respect to the differential pressure. The purpose is to
本発明は、従来のキヤピラリチユーブの流量特
性を改良すべく、第２減圧器として管長さと管直
径の比が2.5〜40となる短管を用いることを特徴
とするものである。
The present invention is characterized in that a short tube having a ratio of tube length to tube diameter of 2.5 to 40 is used as the second pressure reducer in order to improve the flow characteristics of the conventional capillary tube.
第４図は、本発明の一実施例を示す短管の断面
図であり、管長さｌと管内径ｄの比ｌ／ｄの値は
2.5〜40である。
FIG. 4 is a sectional view of a short tube showing an embodiment of the present invention, and the value of the ratio l/d of the tube length l to the tube inner diameter d is
It is between 2.5 and 40.
第５図は、従来のキヤピラリチユーブ（管直径
ｄ＝0.5〜５mm、管長さｌ＝100〜2000mm、ｌ／ｄ
＝100〜1000程度）の流量特性（実線）と、本発
明による短管の流量特性（一点鎖線）の比較を表
わす。図は凝縮温度が低い場合、すなわち冷房運
転で外気温度が低い場合の流量特性であり、ｌ／
ｄを小さくすることにより、最大流量が得られる
差圧ΔPが小さくなつている。従来のキヤピラリ
チユーブでは、差圧ΔPにより流量Ｇが変化する
点（差圧ΔP0，流量G0）でバランスするのに対し
て、本実施例では、このような条件でも差圧ΔP0
で最大流量G′0が得られている。したがつて、最
も差圧ΔPが小さくなる条件（例えば、冷房運転
で外気温度が20℃程度）でも、最大流量が得られ
るように設定された短管を第２減圧器に用いれ
ば、第６図に破線で示されるような作動点が得ら
れ、圧力変動や流量が不足した運転はない。 Figure 5 shows a conventional capillary tube (tube diameter d = 0.5 to 5 mm, tube length l = 100 to 2000 mm, l/d
100 to 1000) (solid line) and the flow rate characteristic (dotted chain line) of the short pipe according to the present invention. The figure shows the flow rate characteristics when the condensing temperature is low, that is, when the outside air temperature is low during cooling operation.
By reducing d, the differential pressure ΔP at which the maximum flow rate is obtained becomes smaller. In conventional capillary tubes, the balance is achieved at the point where the flow rate G changes due to the differential pressure ΔP (differential pressure ΔP 0 , flow rate G 0 ), but in this embodiment, even under these conditions, the differential pressure ΔP 0
The maximum flow rate G′ 0 is obtained at Therefore, if a short pipe set to obtain the maximum flow rate even under conditions where the differential pressure ΔP is the smallest (for example, when the outside temperature is about 20°C during cooling operation) is used for the second pressure reducer, the sixth The operating point shown by the broken line in the figure was obtained, and there was no pressure fluctuation or insufficient flow.
第７図は、第２減圧器のｌ／ｄを変化させて上
記冷房運転（外気温度20℃程度）を行つた場合の
第２減圧器冷媒流量と冷房能力及び成績係数（冷
房能力を電気入力で除した値で運転効率を表わ
す）を表わす。減圧器の抵抗の大きさは、冷房の
標準的な条件（外気温度35℃程度）で最適に設定
されている。第６図に示されるように外気温度が
低下した場合には、ｌ／ｄにより流量Ｇ、能力
Ｑ、成績係数Ｃは大きく変化していることがわか
る。ｌ／ｄが2.5以下では、一般によく知られて
いるオリフイスの流量特性同様常に入口出口の差
圧のみで流量が決定されるため、外気温度の変化
に対して制御性は悪く、この場合は外気温度が低
下しても流量が多く、極度な液戻り運転となり能
力が低下するとともに圧縮機の入力が増加して成
績係数も低下する。一方ｌ／ｄが40以上の場合に
は上述したように流量が不足した運転となり、能
力の低下とともに圧力変動などにより成績係数が
低下する。これに対してｌ／ｄが2.5〜40の範囲
では、安定した適正な流量が得られており、冷房
能力、成績係数ともに最大値が得られている。な
お第４図に示した例は流路断面が円形であるが流
路長さと断面積の比が同程度であれば上述の効果
は得られるものである。 Figure 7 shows the second pressure reducer refrigerant flow rate, cooling capacity, and coefficient of performance (cooling capacity is input electrically The value divided by is the operating efficiency). The resistance of the pressure reducer is optimally set under standard cooling conditions (outside temperature of approximately 35°C). As shown in FIG. 6, it can be seen that when the outside temperature decreases, the flow rate G, capacity Q, and coefficient of performance C change greatly depending on l/d. When l/d is 2.5 or less, the flow rate is always determined only by the differential pressure at the inlet and outlet, similar to the generally well-known flow characteristics of orifices, so controllability is poor in response to changes in outside air temperature. Even when the temperature decreases, the flow rate is large, resulting in extreme liquid return operation, which reduces capacity and increases input to the compressor, reducing the coefficient of performance. On the other hand, when l/d is 40 or more, the flow rate is insufficient as described above, and the coefficient of performance decreases due to pressure fluctuations as well as a decrease in capacity. On the other hand, when l/d is in the range of 2.5 to 40, a stable and appropriate flow rate is obtained, and both the cooling capacity and the coefficient of performance are at their maximum values. In the example shown in FIG. 4, the cross section of the flow path is circular, but the above-mentioned effect can be obtained if the ratio of the flow path length and cross-sectional area is approximately the same.
このように本発明によれば、冷房運転で外気温
度が低下した場合にも、安定した運転点が得られ
ることから、運転効率が向上するとゝもに、圧縮
機への間欠的な液戻りも防止され、信頼性向上が
図られる。
As described above, according to the present invention, a stable operating point can be obtained even when the outside air temperature drops during cooling operation, which improves operating efficiency and prevents intermittent liquid return to the compressor. This will improve reliability.
第１図は、ガスインジエクシヨンサイクルの冷
凍サイクル構成。第２図は、キヤピラリチユーブ
の流量特性。第３図は、ガスインジエクシヨンサ
イクルの第２減圧器にキヤピラリチユーブを用い
た場合の作動点の変化。第４図は、本発明の一実
施例を示す短管の断面図。第５図は、キヤピラリ
チユーブと本発明の短管の流量特性の比較。第６
図は、本発明の短管をガスインジエクシヨンサイ
クルの第２減圧器として作用させた場合の作動点
の変化。第７図は、本発明による特性曲線図であ
る。
１……圧縮機、２……凝縮器、３……第１減圧
器、４……気液分離器、５……第２減圧器、６…
…蒸発器。
Figure 1 shows the refrigeration cycle configuration of the gas injection cycle. Figure 2 shows the flow characteristics of the capillary tube. Figure 3 shows the change in operating point when a capillary tube is used as the second pressure reducer in the gas injection cycle. FIG. 4 is a sectional view of a short tube showing an embodiment of the present invention. FIG. 5 is a comparison of the flow characteristics of the capillary tube and the short tube of the present invention. 6th
The figure shows changes in the operating point when the short tube of the present invention is used as a second pressure reducer in a gas injection cycle. FIG. 7 is a characteristic curve diagram according to the present invention. 1... Compressor, 2... Condenser, 3... First pressure reducer, 4... Gas-liquid separator, 5... Second pressure reducer, 6...
…Evaporator.
Claims (1)
第２減圧器及び蒸発器等を順次接続配管するとと
もに、上記気液分離器に設けられた蒸気流出口と
上記圧縮機に設けられたインジエクシヨンポート
とを連絡するガスインジエクシヨン回路を備えた
ガスインジエクシヨンサイクルにおいて、上記気
液分離器と上記蒸発器との間に設ける第２減圧器
として管長さと管直径の比が2.5〜40となる短管
を用いたことを特徴とする冷凍サイクル。1 Compressor, condenser, first pressure reducer, gas-liquid separator,
A second pressure reducer, an evaporator, etc. are sequentially connected to each other, and a gas injection circuit is provided to connect a vapor outlet provided in the gas-liquid separator with an injection port provided in the compressor. refrigeration, characterized in that in the gas injection cycle, a short tube having a tube length to tube diameter ratio of 2.5 to 40 is used as a second pressure reducer provided between the gas-liquid separator and the evaporator. cycle.
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP10765783A JPS60263A (en) | 1983-06-17 | 1983-06-17 | Refrigeration cycle |
Applications Claiming Priority (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP10765783A JPS60263A (en) | 1983-06-17 | 1983-06-17 | Refrigeration cycle |
Publications (2)
| Publication Number | Publication Date |
|---|---|
| JPS60263A JPS60263A (en) | 1985-01-05 |
| JPH0120692B2 true JPH0120692B2 (en) | 1989-04-18 |
Family
ID=14464723
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| JP10765783A Granted JPS60263A (en) | 1983-06-17 | 1983-06-17 | Refrigeration cycle |
Country Status (1)
| Country | Link |
|---|---|
| JP (1) | JPS60263A (en) |
Families Citing this family (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| US8257319B2 (en) | 2009-06-01 | 2012-09-04 | Sanofi-Aventis Deutschland Gmbh | Drug delivery device inner housing having helical spline |
-
1983
- 1983-06-17 JP JP10765783A patent/JPS60263A/en active Granted
Also Published As
| Publication number | Publication date |
|---|---|
| JPS60263A (en) | 1985-01-05 |
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