以下，參照圖式，對用以實施本發明的形態詳細地進行說明。在說明及圖式中，對相同或相等的構成要素、構件及處理標註相同元件符號，並適當地省略重複說明。所圖示之各部的比例或形狀，為便於說明被便利地設定，若無特別說明則不會被限定地解釋。實施形態為例示，並非對本發明的範圍構成任何限定。實施形態中所記載之所有特徵或其組合，未必為發明的本質者。
圖1係概略表示實施形態之低溫泵系統10之圖。低溫泵系統10具備至少1台低溫泵12和複數台壓縮機14。在此，作為例示性結構而示出低溫泵系統10具有2台低溫泵12和2台壓縮機14之情況。
低溫泵12例如安裝於濺鍍裝置或蒸鍍裝置的真空腔室中，為了將真空腔室內部的真空度提高至所期望的真空程序中所要求之水準而被使用。
低溫泵12具備低溫板16和利用冷媒氣體的隔熱膨脹來冷卻低溫板16之冷凍機18。低溫板16容納於低溫泵12中，在低溫泵12進行動作期間藉由冷凍機18冷卻至極低溫。冷凍機18亦稱為膨脹機或冷頭(cold head)，並且與壓縮機14一同構成極低溫冷凍機。從低溫泵12的進氣口進入之氣體，在被冷卻至極低溫之低溫板16的表面凝結，或者被設置於低溫板16上之吸附材料吸附而被捕集。低溫板16的配置、形狀等低溫泵12的結構能夠適當地採用各種公知的結構，因此在此不做詳述。
壓縮機14以將冷媒氣體供給到冷凍機18的方式並聯連接並同時運轉。壓縮機14構成為，從冷凍機18回收冷媒氣體，並將所回收之冷媒氣體進行升壓，再度將冷媒氣體供給到冷凍機18。
如後所述，在低溫泵系統10進行動作期間，複數台壓縮機14正常狀態下同時運轉。在任一個壓縮機14因某種原因而停止之情況下，剩餘的壓縮機14繼續運轉。
壓縮機14與冷凍機18之間的冷媒氣體的循環是利用冷凍機18內的冷媒氣體的適當之壓力變動與容積變動的組合而進行的，藉此構成產生寒冷之熱力循環，使冷凍機18的冷卻台冷卻至所期望的極低溫。藉此，能夠將熱耦合到冷凍機18的冷卻台之低溫板16冷卻至目標冷卻溫度。冷媒氣體通常為氦氣，但亦可適當使用其他氣體。為便於理解，圖1中用箭頭來表示冷媒氣體的流動方向。
作為一例，冷凍機18為單段式或二段式吉福德-麥克馬洪(Gifford-McMahon；GM)冷凍機，但亦可為脈衝管冷凍機、斯特林冷凍機、或除此以外的類型的極低溫冷凍機。冷凍機18根據極低溫冷凍機的類型具有不同之結構。壓縮機14，不管極低溫冷凍機的類型如何都能夠使用相同的結構。
另外，通常，從壓縮機14供給到冷凍機18之冷媒氣體的壓力和從冷凍機18回收到壓縮機14之冷媒氣體的壓力均明顯高於大氣壓，分別能夠稱為第1高壓及第2高壓。為便於說明，第1高壓及第2高壓亦分別簡稱為高壓及低壓。典型的是，高壓例如在約2～3MPa的範圍，低壓例如在約0.5～1.5MPa的範圍。
壓縮機14具有吐出埠20和吸入埠21。吐出埠20係為了從壓縮機14送出藉由壓縮機14升壓至高壓之冷媒氣體而設置於壓縮機14之冷媒氣體的出口，吸入埠21係為了在壓縮機14接收低壓冷媒氣體而設置於壓縮機14之冷媒氣體的入口。
冷凍機18具有高壓埠22和低壓埠23。高壓埠22係為了在冷凍機18的內部接收高壓動作氣體而設置於冷凍機18之冷媒氣體的入口。低壓埠23係為了從冷凍機18排出藉由冷凍機18內部的冷媒氣體的膨脹而減壓之低壓冷媒氣體而設置於冷凍機18之冷媒氣體的出口。
又，低溫泵系統10具備配管系統24，該配管系統24連接壓縮機14與冷凍機18以使冷媒氣體在它們之間循環。配管系統24具備高壓管線26和低壓管線28。高壓管線26構成為能夠使冷媒氣體從壓縮機14的吐出埠20經過高壓合流部25流到冷凍機18的高壓埠22。低壓管線28構成為能夠使冷媒氣體從冷凍機18的低壓埠23經過低壓合流部27流到壓縮機14的吸入埠21。配管系統24對每個壓縮機14具備吐出側止回閥29和吸入側止回閥30。
高壓管線26具有壓縮機高壓副管線31和冷凍機高壓副管線32。壓縮機高壓副管線31將壓縮機14的吐出埠20連接於高壓合流部25，冷凍機高壓副管線32將冷凍機18的高壓埠22連接於高壓合流部25。高壓管線26係從壓縮機14向冷凍機18的冷媒氣體的流路，因此能夠將從壓縮機14朝向冷凍機18之流動方向稱為高壓管線26的正向，將其相反的方向稱為高壓管線26的反向。正向相當於所圖示之箭頭的方向。吐出側止回閥29在壓縮機高壓副管線31上配置成容許正向的冷媒氣體流，並阻斷反向的冷媒氣體流。
低壓管線28具有壓縮機低壓副管線33和冷凍機低壓副管線34。壓縮機低壓副管線33將壓縮機14的吸入埠21連接於低壓合流部27，冷凍機低壓副管線34將冷凍機18的低壓埠23連接於低壓合流部27。低壓管線28係從冷凍機18向壓縮機14的冷媒氣體的流路，因此能夠將從冷凍機18朝向壓縮機14之流動方向稱為低壓管線28的正向，將其相反的方向稱為低壓管線28的反向。吸入側止回閥30在壓縮機低壓副管線33上配置成容許正向的冷媒氣體流，並阻斷反向的冷媒氣體流。
吐出側止回閥29及吸入側止回閥30皆構成為：在正向上游側(亦即，朝向止回閥的入口側)的冷媒氣體壓力超出正向下游側(亦即，止回閥的出口側)的冷媒氣體壓力之情況下打開，相反地，在正向上游側的冷媒氣體壓力未超出正向下游側的冷媒氣體壓力之情況下關閉。換言之，各止回閥(29、30)當存在流過該止回閥之正向的冷媒氣體流時，止回閥因正向流所引起之壓力損失而自然打開。另一方面，各止回閥(29、30)若在該止回閥的出入口之間產生可能引起冷媒氣體的逆流之壓力差(亦即，出口壓力高於入口壓力)則關閉。如此般利用上游側與下游側之間的壓差作用而開閉之止回閥是一般可獲得的，各止回閥(29、30)能夠適當地採用這種通用的止回閥。
作為一例，高壓管線26及低壓管線28由柔性管構成，但亦可由剛性管構成。又，高壓合流部25及/或低壓合流部27可以構成為單一零件(例如歧管)。可以該單一零件中組裝吐出側止回閥29及/或吸入側止回閥30。
配管系統24具備：設置於吐出側止回閥29的兩側之一組可拆卸的接頭35和設置於吸入側止回閥30的兩側之另一組可拆卸的接頭35。可拆卸的接頭35例如為自封式接頭。
另外，可拆卸的接頭35可以僅設置於吐出側止回閥29的一側(亦即，吐出側止回閥29與吐出埠20之間，或者吐出側止回閥29與高壓合流部25之間)。同樣地，可拆卸的接頭35可以僅設置於吸入側止回閥30的一側。吐出側止回閥29可以一體地組裝於吐出埠20。吸入側止回閥30可以一體地組裝於吸入埠21。
複數個壓縮機14正常狀態下在低溫泵系統10進行動作期間同時運轉。
如圖1所示，在各壓縮機14中經壓縮之高壓冷媒氣體，從壓縮機14的吐出埠20送出到壓縮機高壓副管線31。由於冷媒氣體向高壓管線26的正向流動，因此能夠通過吐出側止回閥29進行流動。來自複數個壓縮機14的冷媒氣體流在高壓合流部25中一度合流，並再度分流到冷凍機高壓副管線32。冷媒氣體從冷凍機高壓副管線32經過冷凍機18的高壓埠22供給到冷凍機18。如此，低溫泵系統10能夠將高壓冷媒氣體從複數個壓縮機14供給到低溫泵12。
從各冷凍機18排出之低壓冷媒氣體，從冷凍機18的低壓埠23流過冷凍機低壓副管線34。冷媒氣體流在低壓合流部27中一度合流，並再度分流到壓縮機低壓副管線33。由於冷媒氣體向低壓管線28的正向流動，因此能夠通過吸入側止回閥30進行流動。冷媒氣體從壓縮機低壓副管線33經過壓縮機14的吸入埠21被壓縮機14回收。如此，低溫泵系統10能夠將低壓冷媒氣體從低溫泵12往複數個壓縮機14進行回收。
圖2係概略表示實施形態之低溫泵系統10中之冷媒氣體的流動之圖。與圖1所示之低溫泵系統10的正常動作不同，圖2中示出複數台壓縮機14中的某一壓縮機14因某種原因而停止之異常時冷媒氣體的流動。由於低溫泵系統10本身無法控制或難以處理之各種外部原因，例如停電、冷卻設備的故障，或者氣溫或濕度、氣壓等周圍環境的異常變動等，可能使壓縮機14異常停止。
在某一壓縮機14異常停止之情況下，剩餘的壓縮機14繼續運轉，藉此能夠避免低溫泵系統10的停止。圖2中示出低溫泵系統10具有2台壓縮機14之例子，因此一個壓縮機14停止，另一個壓縮機14繼續正常運轉。
為便於說明，將停止中之壓縮機稱為第1壓縮機14a，將動作中之壓縮機稱為第2壓縮機14b。又，將附隨於第1壓縮機14a之吐出側止回閥29及吸入側止回閥30分別稱為第1吐出側止回閥29a及第1吸入側止回閥30a。同樣地，將附隨於第2壓縮機14b之吐出側止回閥29及吸入側止回閥30分別稱為第2吐出側止回閥29b及第2吸入側止回閥30b。
該情況下，如圖2所示，由第1壓縮機14a壓縮後之高壓冷媒氣體從第1壓縮機14a的吐出埠20送出到高壓管線26。冷媒氣體從壓縮機高壓副管線31經過高壓合流部25向冷凍機高壓副管線32進行分歧，並流入各冷凍機18的高壓埠22。由於冷媒氣體向高壓管線26的正向流動，能夠通過第1吐出側止回閥29a進行流動。
另一方面，由於第2壓縮機14b停止，因此冷媒氣體不會從第2壓縮機14b的吐出埠20吐出。因此，關於第2吐出側止回閥29b，其正向上游側的冷媒氣體壓力低於正向下游側的冷媒氣體壓力，第2吐出側止回閥29b關閉。因此，第2吐出側止回閥29b阻斷通過了壓縮機高壓副管線31之向第2壓縮機14b的冷媒氣體的逆流。
如此，能夠通過高壓管線26將高壓冷媒氣體從第1壓縮機14a供給到冷凍機18。又，可防止從第1壓縮機14a向第2壓縮機14的通過了高壓管線26之逆流。
從冷凍機18排出之低壓冷媒氣體從冷凍機18的低壓埠23送出到低壓管線28。冷媒氣體從冷凍機低壓副管線34經過低壓合流部27及壓縮機低壓副管線33流入到第1壓縮機14a的吸入埠21。由於冷媒氣體向低壓管線28的正向流動，能夠通過第1吸入側止回閥30a進行流動。
另一方面，由於第2壓縮機14b停止，因此冷媒氣體不會從第2壓縮機14b的吸入埠21被吸入。因此，關於第2吸入側止回閥30b，在其正向下游側壓力變得比正向上游側更高，第2吸入側止回閥30b關閉。因此，第2吸入側止回閥30b阻斷通過了壓縮機低壓副管線33之向第2壓縮機14b的冷媒氣體的逆流。
如此，能夠通過低壓管線28將低壓冷媒氣體從冷凍機18往第1壓縮機14a進行回收。又，可防止從第2壓縮機14向第1壓縮機14a的通過了低壓管線28之逆流。
另外，在壓縮機14中，通常，當停止時吐出埠20和吸入埠21被均壓化。亦即，吐出埠20和吸入埠21均成為高壓和低壓的平均壓力(例如，若高壓為2MPa，低壓為0.6 MPa，則平均壓力為1.3MPa)。因此，第2吐出側止回閥29b和第2吸入側止回閥30b都是出口壓力明顯高於入口壓力，藉由該壓力差被確實地關閉。
圖3係與實施形態之低溫泵系統10有關之控制方塊圖。圖3示出低溫泵系統10的相關部分，關於複數個低溫泵12中的1個示出內部的細節，關於其他低溫泵12，因相同而省略圖示。同樣地，關於複數個壓縮機14示出細節，其他壓縮機14與其相同，因此省略內部的圖示。
另外，這種低溫泵系統10的控制結構，在硬體結構上能用以電腦的CPU或記憶體為首之元件或電路來實現，在軟體結構上能用電腦程式等來實現，在圖3中適當地繪製為藉由其等的協作來實現之功能方塊。本領域技術人員應可理解，該等功能方塊能夠藉由硬體、軟體的組合以各種形式來實現。
低溫泵系統10具備低溫泵控制器(以下，亦稱為CP控制器)100。CP控制器100控制低溫泵12(亦即，冷凍機18)及壓縮機14。CP控制器100具備：執行各種運算處理之CPU、儲存各種控制程式之ROM、作為用以儲存資料或執行程式的工作區域而被利用之RAM、輸入/輸出介面、記憶體等。又，CP控制器100構成為亦能夠與上位控制器(未圖示)進行通訊，該上位控制器用以控制安裝了低溫泵12之真空程序裝置。
CP控制器100是與低溫泵12及壓縮機14分開設置。CP控制器100與低溫泵12及壓縮機14連接成能夠互相通訊。低溫泵12分別具備：用以與CP控制器100通訊之進行輸入輸出處理的IO模組50。另外，CP控制器100亦可以一體地搭載於任一個低溫泵12或壓縮機14。
如上所述，CP控制器100以能夠通訊之方式連接於各低溫泵12的IO模組50。IO模組50包括冷凍機變頻器52及訊號處理部54。冷凍機18具備作為驅動冷凍機18的熱力循環之驅動源的冷凍機馬達56。冷凍機變頻器52調整從外部電源例如商用電源供給之既定的電壓及頻率的電力，並供給到冷凍機馬達56。由CP控制器100來控制應供給到冷凍機馬達56之電壓及頻率。低溫泵12具備低溫板溫度感測器58。低溫板溫度感測器58測量冷凍機18的冷卻台及/或低溫板16(參閱圖1)的溫度。
CP控制器100依據感測器輸出訊號來決定指令控制量。訊號處理部54將從CP控制器100發送之指令控制量中繼到冷凍機變頻器52。例如，訊號處理部54將來自CP控制器100的指令訊號轉換成能夠由冷凍機變頻器52處理之訊號，並發送到冷凍機變頻器52。指令訊號包括表示冷凍機18的運轉頻率之訊號。又，訊號處理部54將低溫泵12的各種感測器的輸出中繼到CP控制器100。例如，訊號處理部54將感測器輸出訊號轉換成能夠由CP控制器100處理之訊號，並發送到CP控制器100。
在IO模組50的訊號處理部54上連接著包括低溫板溫度感測器58之各種感測器。低溫板溫度感測器58週期性地測量低溫板16的溫度，並輸出表示測量溫度值之訊號。低溫板溫度感測器58的測量溫度訊號每隔既定時間輸入到CP控制器100，溫度測量值儲存並保持在CP控制器100的既定的儲存區域中。
冷凍機18的運轉頻率(亦稱為運轉速度)表示冷凍機馬達56的運轉頻率或轉速、冷凍機變頻器52的運轉頻率、冷凍機18的熱力循環(例如GM循環等的冷凍循環)的頻率或其中任一個。熱力循環的頻率為在冷凍機18中進行之熱力循環的每單位時間的次數。
CP控制器100構成為，以目標冷卻溫度與低溫板的測量溫度的偏差的函數的形式(例如藉由PID控制)決定冷凍機18的運轉頻率。CP控制器100將所決定之運轉頻率輸出到冷凍機變頻器52。冷凍機變頻器52將輸入電力轉換成具有從CP控制器100輸入之運轉頻率。對冷凍機變頻器52的輸入電力是從冷凍機電源(未圖示)供給。冷凍機變頻器52將轉換後之電力輸出到冷凍機馬達56。如此，冷凍機馬達56以由CP控制器100所決定並從冷凍機變頻器52輸出之運轉頻率而被驅動。
當對低溫泵12的熱負載增加時，低溫板16的溫度有可能提高。在低溫板溫度感測器58的測量溫度高於目標溫度之情況下，CP控制器100使冷凍機18的運轉頻率增加。其結果，冷凍機18中之熱力循環的頻率亦增加(亦即，冷凍機18的冷凍能力提高)，低溫板16朝向目標溫度被冷卻。相反地，在低溫板溫度感測器58的測量溫度低於目標溫度之情況下，冷凍機18的運轉頻率減少，冷凍能力降低，低溫板16朝向目標溫度升溫。如此，能夠將低溫板16的溫度保持在目標溫度附近的溫度範圍內。由於能夠根據熱負載適當地調整冷凍機18的運轉頻率，因此這種控制有助於減小低溫泵12的耗電。
如此，在冷凍機18中的低溫板冷卻中所使用之冷媒氣體流量，以因應對低溫泵12的熱負載而將低溫板16的溫度維持在目標溫度的方式發生變化。假使從壓縮機14供給到冷凍機18之冷媒氣體流量不足，即使冷凍機18的運轉頻率增加，冷凍機18的冷凍能力亦不會充分地增加。因此，實施形態之低溫泵系統10構成為使來自壓縮機14的冷媒氣體吐出流量成為可變。以下，對如此般壓縮機控制的一例進行說明。
又，壓縮機14具備壓縮機控制器60、壓縮機變頻器62、壓縮機馬達64、第1壓力感測器66及第2壓力感測器68。壓縮機14例如構成為渦卷方式、旋轉式或使冷媒氣體升壓之其他泵，壓縮機馬達64作為驅動壓縮機14之驅動源而被設置。
與冷凍機18的情況同樣地，壓縮機控制器60決定壓縮機14的運轉頻率，並將所決定之運轉頻率輸出到壓縮機變頻器62。壓縮機變頻器62按照從壓縮機控制器60輸入之運轉頻率而轉換輸入電力，並將轉換後之電力輸出到壓縮機馬達64。如此，壓縮機馬達64以從由壓縮機控制器60決定並從壓縮機變頻器62輸出之運轉頻率而被驅動。在此，壓縮機14的運轉頻率例如係指壓縮機變頻器62的運轉頻率、壓縮機馬達64的運轉頻率或轉速。
在壓縮機14的內部，將第1壓力感測器66設置成測量低溫泵系統10的高壓(例如高壓管線26的壓力)，將第2壓力感測器68設置成測量低溫泵系統10的低壓(例如低壓管線28的壓力)。第1壓力感測器66及第2壓力感測器68分別週期性地測量冷媒氣體的壓力，並將表示測量壓力值之訊號輸出到壓縮機控制器60。壓縮機控制器60將測量壓力訊號及/或壓縮機14的運轉頻率發送到CP控制器100亦可。
壓縮機控制器60構成為依據第1壓力感測器66及/或第2壓力感測器68的測量壓力而控制壓縮機14的運轉頻率。例如壓縮機控制器60構成為，以壓縮機14的吐出側和吸入側的壓差與目標壓差的偏差的函數的形式(例如藉由PID控制)決定壓縮機14的運轉頻率。如此的壓縮機14的控制有時被稱為“壓差恆定控制”。另外，根據需要，壓差的目標值也可以在壓差恆定控制的執行中變更。在壓差恆定控制中，壓縮機控制器60求出第1壓力感測器66的測量壓力與第2壓力感測器68的測量壓力的壓差。壓縮機控制器60以使該測量壓差與壓差目標值一致之方式決定壓縮機14的運轉頻率。若測量壓差大於壓差目標值，則壓縮機控制器60使運轉頻率降低，若測量壓差小於壓差目標值，則使運轉頻率增加。
在藉由冷凍機18的低溫板冷卻中所使用之冷媒氣體流量與冷凍機18的運轉頻率成正比，例如由冷凍機18的內部容積和冷凍機18的運轉頻率的乘積來求出。冷凍機18的運轉頻率愈增加，從壓縮機14應供給到冷凍機18之冷媒氣體流量愈增加。此時，若壓縮機14的運轉頻率低而使來自壓縮機14的冷媒氣體供給不足，則壓縮機14的吐出側的壓力降低。當冷凍機18的運轉頻率增加時，從冷凍機18往壓縮機14應回收之冷媒氣體流量亦增加。此時，若壓縮機14的運轉頻率低，則壓縮機14不會充分地回收從冷凍機18排出之冷媒氣體，因此壓縮機14的吸入側的壓力提高。如此，冷凍機18的運轉頻率的增加導致壓縮機14的吐出側與吸入側的壓差趨於減小。相反地，冷凍機18的運轉頻率的減少導致壓縮機14的吐出側與吸入側的壓差趨於增大。
依據壓縮機14的壓差恆定控制，當對低溫泵12的負載增加而使冷凍機18的運轉頻率增加時，壓縮機14的運轉頻率增加以抑制壓縮機14的吐出側與吸入側的壓差降低，從壓縮機14向冷凍機18的冷媒氣體供給亦增加。另一方面，當對低溫泵12的負載降低而使冷凍機18的運轉頻率減少時，壓縮機14的運轉頻率減少，從壓縮機14向冷凍機18的冷媒氣體供給亦被抑制。由於能夠根據對低溫泵系統10的負載適當地調整壓縮機14的運轉頻率，因此壓差恆定控制有助於減少低溫泵系統10的耗電。
然而，典型的低溫泵系統能夠構成為僅具有1台壓縮機。相對於此，在實施形態之低溫泵系統10中，壓縮機14不僅1台，而且還追加了另1台壓縮機14。低溫泵系統10關於壓縮機14具有冗餘度。這2台壓縮機14作為對低溫泵12的冷媒氣體供給源而在低溫泵系統10進行動作期間同時運轉。
2台壓縮機14的冷媒氣體供給能力的總計設定為不低於低溫泵12的各冷凍機18執行低溫板冷卻所需冷媒氣體流量的總計。在此，壓縮機14的冷媒氣體供給能力，係指例如在壓縮機14以最大的運轉頻率運轉時實現之壓縮機14的最大吐出流量。冷凍機18所需冷媒氣體流量，係指例如在冷凍機18以最大的運轉頻率運轉時使用於冷凍機18中之冷媒氣體流量。因此，當將2台壓縮機14的冷媒氣體供給能力用Qc1
、Qc2
表示、且將2台冷凍機18所需冷媒氣體流量用qr1
、qr2
表示時，以下關係成立。
Qc1
+Qc2
≥qr1
+qr2
藉由如此般設定壓縮機14的冷媒氣體供給能力，能夠將冷媒氣體藉由2台壓縮機14的同時運轉充分地供給到2台冷凍機18。由於能夠避免冷凍機18中之冷媒氣體的不足，因此能夠將低溫板16維持在目標溫度，並能夠繼續進行低溫泵系統10的運轉。
而且，在實施形態之低溫泵系統10中，關於2台壓縮機14中的任何壓縮機14，該壓縮機14的冷媒氣體供給能力設定為不低於各低溫泵12的冷凍機18執行低溫板冷卻所需冷媒氣體流量的總計。亦即，低溫泵系統10亦滿足以下關係。
Qc1
≥qr1
+qr2
且Qc2
≥qr1
+qr2
如參閱圖2已說明，假定一個壓縮機14因某種原因而停止之狀況。然而，低溫泵系統10能夠藉由未停止之另一個壓縮機14將冷媒氣體充分地供給到2台冷凍機18。如此，即使在1台壓縮機14不運轉之狀況下，低溫泵系統10亦能夠將各低溫泵12的低溫板16維持在目標溫度，並能夠繼續進行低溫泵系統10的運轉。
實施形態之低溫泵系統10的結構能夠如下般通則化。低溫泵系統10具備M台低溫泵12和N+1台壓縮機14，該N+1台壓縮機14以將冷媒氣體供給到各低溫泵12的冷凍機18的方式並聯連接並同時運轉。其中，M、N分別為正整數。作為一例，正整數M例如可以為1或比其更大，2或比其更大，3或比其更大，5或比其更大，或者10或比其更大。正整數M例如可以為20或比其更小，10或比其更小，5或比其更小，或者3或比其更小。正整數N例如可以為1或比其更大，2或比其更大，3或比其更大，5或比其更大，或者10或比其更大。正整數N例如可以為20或比其更小，10或比其更小，5或比其更小，或者3或比其更小。
關於N+1台壓縮機14中的任何N台壓縮機14，該N台壓縮機14的冷媒氣體供給能力的總計設定為不低於各低溫泵12的冷凍機18執行低溫板冷卻所需冷媒氣體流量的總計。因此，當將N+1台壓縮機14的冷媒氣體供給能力用Qc1
、Qc2
、……、QcN
、QcN+1
表示，將M台冷凍機18的必要冷媒氣體流量用qr1
、qr2
、……、qrM
表示時，低溫泵系統10滿足以下所有關係。
ΣQc-Qc1
≥Σqr
ΣQc-Qc2
≥Σqr
……
ΣQc-QcN
≥Σqr
ΣQc-QcN+1
≥Σqr，
在此，ΣQc=Qc1
+Qc2
+……+QcN
+QcN+1
(亦即，N+1台壓縮機14的冷媒氣體供給能力的總計)、Σqr=qr1
+qr2
+……+qrM
(亦即，M台冷凍機18的必要冷媒氣體流量的總計)。因此，上述各式的左邊表示N+1台壓縮機14中的任意的N台壓縮機14的冷媒氣體供給能力的總計。
如此一來，即使任一個壓縮機14因某種原因而停止，低溫泵系統10亦能夠藉由未停止之剩餘的壓縮機14來將冷媒氣體充分地供給到M台低溫泵12的冷凍機18。低溫泵系統10在任一個壓縮機14停止期間，亦能夠將各低溫泵12的低溫板16維持在目標溫度，並能夠繼續進行低溫泵系統10的運轉。
又，在N+1台壓縮機14全部運轉之正常狀況下，低溫泵系統10包括1台剩餘的壓縮機14。因此，與低溫泵系統10僅包括N台壓縮機14之情況相比，N+1台壓縮機14的每一個應供給之冷媒氣體流量更少即可。因此，藉由實施形態之低溫泵系統10，能夠以較低的負載(亦即，運轉頻率)來運轉各壓縮機14，這有助於延長壓縮機14的壽命。
又，低溫泵系統10具備控制N+1台壓縮機14之控制部(例如，壓縮機控制器60或CP控制器100)。控制部構成為將各壓縮機14控制成：當同時運轉之壓縮機14的數量從N+1台減少到N台時，使同時運轉之N台壓縮機14各自的冷媒氣體供給增加。
適合於如此的壓縮機控制之一例為上述壓差恆定控制。若考慮到複數台壓縮機14中的任一個壓縮機14停止之狀況，因為冷媒氣體供給流量的總計降低相當於停止中之一台壓縮機14的量，可能使高壓管線26的壓力降低且低壓管線28的壓力提高。亦即，若任一個壓縮機14停止，則剩餘的壓縮機14的每一個的吐出側與吸入側的壓差趨於降低。依據壓差恆定控制，各壓縮機14的運轉頻率增加，以使該種壓差的降低恢復到目標壓差。如此，低溫泵系統10能夠將各壓縮機14控制成：當同時運轉之壓縮機14的數量從N+1台減少到N台時，使同時運轉之N台壓縮機14各自的冷媒氣體供給增加。
再者，低溫泵系統10的配管系統24是對每一個壓縮機14具備吐出側止回閥29和吸入側止回閥30。如此一來，即使複數台壓縮機14中的任一個壓縮機14停止，仍能夠防止從運轉中之剩餘的壓縮機14向停止中的壓縮機14的冷媒氣體的逆流。由於吐出側止回閥29和吸入側止回閥30藉由壓差機械地關閉，因此不需要電控制便能夠從低溫泵系統10將停止中之壓縮機14自然地斷開。
尤其，吐出側止回閥29和吸入側止回閥30能夠採用藉由出入口之間的壓差來進行動作之通用的止回閥，該等止回閥具有相對簡單之結構且價格低廉。與將用於斷開的電控制閥設置於配管系統24相比，能夠更簡單地構成配管系統24，這會有助於降低低溫泵系統10的製造成本。另外，根據需要，配管系統24可以具備控制閥來代替吐出側止回閥29及/或吸入側止回閥30，該控制閥構成為阻斷向停止中之壓縮機14的冷媒氣體逆流。
又，配管系統24具備設置於吐出側止回閥29的兩側之一組可拆卸的接頭35。配管系統24具備設置於吸入側止回閥30的兩側之另一組可拆卸的接頭35。如此一來，作業人員能夠從低溫泵系統10拆除已停止之壓縮機14並實施維護。或者，作業人員能夠從低溫泵系統10拆除壓縮機14，並與新的壓縮機或完成維護之其他壓縮機進行更換。由於能夠一邊繼續進行低溫泵系統10的運轉，一邊進行如此的維護作業，因此方便。
以上，依據實施例對本發明進行了說明。本發明並不限定於上述實施形態而能夠進行各種設計變更，本領域技術人員當然能夠理解，能夠進行各種變形例，又，該等變形例亦包括在本發明的範圍內。又，與一實施形態相關聯地已說明之各種特徵，亦能夠應用到其他實施形態中。根據組合而生成之新的實施形態同時具有被組合之實施形態各自的效果。
在上述實施形態中，冷凍機18所需冷媒氣體流量係指例如當冷凍機18以最大的運轉頻率運轉時使用於冷凍機18之冷媒氣體流量。實際上，要求冷凍機18的最大運轉頻率之情況，限定於低溫泵系統10起動時(此時，希望冷凍機18從室溫高速冷卻到極低溫)等而是罕見的。如此，在低溫泵系統10起動並穩定地運行的狀態下，冷凍機18所需冷媒氣體流量並不需要很多。於是，冷凍機18所需冷媒氣體流量亦可以指當冷凍機18以某一運轉頻率臨限值運轉時使用於冷凍機18之冷媒氣體流量。該運轉頻率臨限值小於最大運轉頻率。如此一來，能夠將壓縮機14的冷媒氣體供給能力設計得更低，因此可謀求各壓縮機14的小型化、低溫泵系統10的製造成本的降低。
低溫泵系統10亦可以具備至少一台低溫泵12和同時運轉之比N+1台更多的壓縮機14(例如N+2台或N+3台壓縮機14)。關於比N+1台更多的壓縮機14中的任何N台壓縮機14，該N台壓縮機14的冷媒氣體供給能力的總計設定為不低於至少一台低溫泵12的各冷凍機18執行低溫板冷卻所需冷媒氣體流量的總計。如此一來，低溫泵系統10關於壓縮機14進一步冗餘化，例如即使2台或3台壓縮機14已停止，亦能夠繼續進行低溫泵系統10的運轉。
或者，超出N+1台之剩餘的壓縮機14，作為正常狀態下不與其他壓縮機14同時運轉之備用壓縮機而設置於低溫泵系統10亦可。Hereinafter, referring to the drawings, a form for implementing the present invention will be described in detail. In the description and the drawings, the same or equivalent constituent elements, members, and processes are denoted by the same element symbols, and repetitive descriptions are omitted as appropriate. The proportions or shapes of the illustrated parts are conveniently set for convenience of explanation, and will not be limitedly interpreted unless otherwise specified. The embodiments are illustrative and do not limit the scope of the present invention. All the features or combinations described in the embodiments are not necessarily the essence of the invention. FIG. 1 is a diagram schematically showing a cryopump system 10 of an embodiment. The cryopump system 10 includes at least one cryopump 12 and a plurality of compressors 14. Here, as an exemplary structure, a case where the cryopump system 10 has two cryopumps 12 and two compressors 14 is shown. The cryopump 12 is installed in a vacuum chamber of a sputtering apparatus or an evaporation apparatus, for example, and is used to increase the vacuum degree inside the vacuum chamber to a level required in a desired vacuum procedure. The cryopump 12 includes a cryopanel 16 and a refrigerator 18 that cools the cryopanel 16 by adiabatic expansion of refrigerant gas. The cryopanel 16 is accommodated in the cryopump 12 and is cooled to an extremely low temperature by the refrigerator 18 during the operation of the cryopump 12. The freezer 18 is also called an expander or cold head, and together with the compressor 14 constitutes a very low temperature freezer. The gas entering from the intake port of the cryopump 12 is condensed on the surface of the cryopanel 16 cooled to an extremely low temperature, or is adsorbed by the adsorbent provided on the cryopanel 16 and trapped. Various configurations of the cryopump 12 such as the arrangement and shape of the cryopanel 16 can be appropriately adopted, and therefore will not be described in detail here. The compressor 14 is connected in parallel so as to supply refrigerant gas to the refrigerator 18, and operates simultaneously. The compressor 14 is configured to recover the refrigerant gas from the refrigerator 18 and boost the recovered refrigerant gas to supply the refrigerant gas to the refrigerator 18 again. As will be described later, during the operation of the cryopump system 10, a plurality of compressors 14 simultaneously operate in a normal state. When any compressor 14 is stopped for some reason, the remaining compressors 14 continue to operate. The circulation of the refrigerant gas between the compressor 14 and the refrigerator 18 is performed by using a combination of appropriate pressure fluctuations and volume fluctuations of the refrigerant gas in the refrigerator 18, thereby forming a thermal cycle that generates cold, so that the refrigerator 18 The cooling table is cooled to the desired extremely low temperature. Thereby, the low temperature plate 16 thermally coupled to the cooling stage of the refrigerator 18 can be cooled to the target cooling temperature. The refrigerant gas is usually helium, but other gases can be used as appropriate. For ease of understanding, the flow direction of the refrigerant gas is indicated by arrows in FIG. 1. As an example, the freezer 18 is a single-stage or two-stage Gifford-McMahon (GM) freezer, but it may also be a pulse tube freezer, a Stirling freezer, or otherwise Type of extremely low temperature freezer. The refrigerator 18 has a different structure according to the type of the extremely low temperature refrigerator. The compressor 14 can use the same structure regardless of the type of the cryogenic refrigerator. In addition, in general, the pressure of the refrigerant gas supplied from the compressor 14 to the refrigerator 18 and the pressure of the refrigerant gas recovered from the refrigerator 18 to the compressor 14 are significantly higher than atmospheric pressure, and can be referred to as the first high pressure and the second high pressure, respectively . For ease of explanation, the first high pressure and the second high pressure are also referred to as high pressure and low pressure, respectively. Typically, the high pressure is in the range of about 2 to 3 MPa, and the low pressure is in the range of about 0.5 to 1.5 MPa, for example. The compressor 14 has a discharge port 20 and a suction port 21. The discharge port 20 is a refrigerant gas outlet provided in the compressor 14 for sending refrigerant gas boosted by the compressor 14 to a high pressure from the compressor 14, and the suction port 21 is provided in the compressor 14 for receiving low-pressure refrigerant gas The refrigerant gas inlet of the compressor 14. The refrigerator 18 has a high-pressure port 22 and a low-pressure port 23. The high-pressure port 22 is a refrigerant gas inlet provided in the refrigerator 18 in order to receive a high-pressure operating gas inside the refrigerator 18. The low-pressure port 23 is a refrigerant gas outlet provided in the refrigerator 18 to discharge the low-pressure refrigerant gas decompressed by the expansion of the refrigerant gas inside the refrigerator 18 from the refrigerator 18. In addition, the cryopump system 10 includes a piping system 24 that connects the compressor 14 and the refrigerator 18 to circulate refrigerant gas therebetween. The piping system 24 includes a high-pressure line 26 and a low-pressure line 28. The high-pressure line 26 is configured to allow refrigerant gas to flow from the discharge port 20 of the compressor 14 to the high-pressure port 22 of the refrigerator 18 through the high-pressure junction 25. The low-pressure line 28 is configured to allow refrigerant gas to flow from the low-pressure port 23 of the refrigerator 18 through the low-pressure junction 27 to the suction port 21 of the compressor 14. The piping system 24 includes a discharge-side check valve 29 and a suction-side check valve 30 for each compressor 14. The high-pressure line 26 has a compressor high-pressure auxiliary line 31 and a refrigerator high-pressure auxiliary line 32. The compressor high-pressure auxiliary line 31 connects the discharge port 20 of the compressor 14 to the high-pressure junction 25, and the refrigerator high-pressure auxiliary line 32 connects the high-pressure port 22 of the refrigerator 18 to the high-pressure junction 25. The high-pressure line 26 is a refrigerant gas flow path from the compressor 14 to the refrigerator 18, so the direction of the flow from the compressor 14 to the refrigerator 18 can be referred to as the positive direction of the high-pressure line 26, and the opposite direction can be referred to as high pressure. Reverse of line 26. The positive direction corresponds to the direction of the arrow shown. The discharge-side check valve 29 is arranged on the compressor high-pressure auxiliary line 31 to allow the forward refrigerant gas flow and block the reverse refrigerant gas flow. The low-pressure line 28 has a compressor low-pressure auxiliary line 33 and a refrigerator low-pressure auxiliary line 34. The compressor low-pressure sub-line 33 connects the suction port 21 of the compressor 14 to the low-pressure junction 27, and the refrigerator low-pressure sub-line 34 connects the low-pressure port 23 of the refrigerator 18 to the low-pressure junction 27. The low-pressure line 28 is a refrigerant gas flow path from the refrigerator 18 to the compressor 14. Therefore, the flow direction from the refrigerator 18 to the compressor 14 can be referred to as the forward direction of the low-pressure line 28, and the opposite direction can be referred to as low pressure. Reverse of pipeline 28. The suction side check valve 30 is arranged on the compressor low-pressure sub-line 33 to allow the forward refrigerant gas flow and block the reverse refrigerant gas flow. The discharge-side check valve 29 and the suction-side check valve 30 are both configured such that the refrigerant gas pressure on the positive upstream side (that is, toward the inlet side of the check valve) exceeds the positive downstream side (that is, the check valve (On the outlet side), the refrigerant gas pressure is opened, and conversely, the refrigerant gas pressure on the forward upstream side does not exceed the refrigerant gas pressure on the forward downstream side. In other words, when there is a flow of refrigerant gas flowing through the check valve in the forward direction, the check valve naturally opens due to the pressure loss caused by the forward flow. On the other hand, each check valve (29, 30) closes if there is a pressure difference between the check valve's inlet and outlet that may cause a reverse flow of refrigerant gas (that is, the outlet pressure is higher than the inlet pressure). Check valves that open and close by utilizing the pressure difference between the upstream side and the downstream side are generally available, and each check valve (29, 30) can appropriately adopt such a common check valve. As an example, the high-pressure line 26 and the low-pressure line 28 are composed of flexible pipes, but they may be composed of rigid pipes. In addition, the high-pressure junction 25 and/or the low-pressure junction 27 may be configured as a single component (for example, a manifold). The discharge side check valve 29 and/or the suction side check valve 30 can be assembled in this single part. The piping system 24 includes one set of detachable joints 35 provided on both sides of the discharge side check valve 29 and another set of detachable joints 35 provided on both sides of the suction side check valve 30. The detachable joint 35 is, for example, a self-sealing joint. In addition, the detachable joint 35 may be provided only on one side of the discharge-side check valve 29 (that is, between the discharge-side check valve 29 and the discharge port 20, or between the discharge-side check valve 29 and the high-pressure junction 25 between). Similarly, the detachable joint 35 may be provided only on one side of the suction side check valve 30. The discharge side check valve 29 can be integrally assembled to the discharge port 20. The suction side check valve 30 can be integrally assembled to the suction port 21. The plurality of compressors 14 are simultaneously operated during the operation of the cryopump system 10 in the normal state. As shown in FIG. 1, the high-pressure refrigerant gas compressed in each compressor 14 is sent from the discharge port 20 of the compressor 14 to the compressor high-pressure auxiliary line 31. Since the refrigerant gas flows in the forward direction of the high-pressure line 26, it can flow through the discharge-side check valve 29. The refrigerant gas flows from the plurality of compressors 14 are merged once in the high-pressure merge section 25, and are again diverted to the high-pressure auxiliary line 32 of the refrigerator. The refrigerant gas is supplied to the refrigerator 18 from the refrigerator high-pressure sub-line 32 through the high-pressure port 22 of the refrigerator 18. In this way, the cryopump system 10 can supply high-pressure refrigerant gas from the plurality of compressors 14 to the cryopump 12. The low-pressure refrigerant gas discharged from each refrigerator 18 flows from the low-pressure port 23 of the refrigerator 18 through the low-pressure sub-line 34 of the refrigerator. The refrigerant gas flow is merged once in the low-pressure merge section 27, and is again divided into the compressor low-pressure auxiliary line 33. Since the refrigerant gas flows in the forward direction of the low-pressure line 28, it can flow through the suction side check valve 30. The refrigerant gas is recovered by the compressor 14 from the compressor low-pressure auxiliary line 33 through the suction port 21 of the compressor 14. In this way, the cryopump system 10 can recover the low-pressure refrigerant gas from the cryopump 12 to and from the compressor 14. 2 is a diagram schematically showing the flow of refrigerant gas in the cryopump system 10 of the embodiment. Unlike the normal operation of the cryopump system 10 shown in FIG. 1, FIG. 2 shows the flow of refrigerant gas during an abnormality when one of the compressors 14 stops due to some reason. The compressor 14 may be abnormally stopped due to various external reasons that the cryopump system 10 itself cannot control or is difficult to deal with, such as power failure, failure of cooling equipment, or abnormal changes in the surrounding environment such as air temperature, humidity, and air pressure. When a certain compressor 14 stops abnormally, the remaining compressors 14 continue to operate, whereby the stop of the cryopump system 10 can be avoided. FIG. 2 shows an example in which the cryopump system 10 has two compressors 14, so one compressor 14 stops and the other compressor 14 continues to operate normally. For convenience of description, the compressor in stop is referred to as the first compressor 14a, and the compressor in operation is referred to as the second compressor 14b. In addition, the discharge-side check valve 29 and the suction-side check valve 30 accompanying the first compressor 14a are referred to as a first discharge-side check valve 29a and a first suction-side check valve 30a, respectively. Similarly, the discharge-side check valve 29 and the suction-side check valve 30 accompanying the second compressor 14b are referred to as a second discharge-side check valve 29b and a second suction-side check valve 30b, respectively. In this case, as shown in FIG. 2, the high-pressure refrigerant gas compressed by the first compressor 14 a is sent from the discharge port 20 of the first compressor 14 a to the high-pressure line 26. The refrigerant gas branches from the compressor high-pressure sub-line 31 through the high-pressure junction 25 to the refrigerator high-pressure sub-line 32 and flows into the high-pressure port 22 of each refrigerator 18. Since the refrigerant gas flows forward to the high-pressure line 26, it can flow through the first discharge side check valve 29a. On the other hand, since the second compressor 14b is stopped, the refrigerant gas is not discharged from the discharge port 20 of the second compressor 14b. Therefore, regarding the second discharge side check valve 29b, the refrigerant gas pressure on the forward upstream side is lower than the refrigerant gas pressure on the forward downstream side, and the second discharge side check valve 29b is closed. Therefore, the second discharge side check valve 29b blocks the reverse flow of the refrigerant gas that has passed through the compressor high-pressure sub-line 31 to the second compressor 14b. In this way, the high-pressure refrigerant gas can be supplied from the first compressor 14 a to the refrigerator 18 through the high-pressure line 26. In addition, it is possible to prevent backflow from the first compressor 14a to the second compressor 14 through the high-pressure line 26. The low-pressure refrigerant gas discharged from the refrigerator 18 is sent to the low-pressure line 28 from the low-pressure port 23 of the refrigerator 18. The refrigerant gas flows into the suction port 21 of the first compressor 14a from the low-pressure sub-line 34 of the refrigerator through the low-pressure junction 27 and the low-pressure sub-line 33 of the compressor. Since the refrigerant gas flows forward to the low-pressure line 28, it can flow through the first suction side check valve 30a. On the other hand, since the second compressor 14b is stopped, the refrigerant gas is not sucked from the suction port 21 of the second compressor 14b. Therefore, with respect to the second suction side check valve 30b, the pressure on the positive downstream side becomes higher than that on the forward upstream side, and the second suction side check valve 30b is closed. Therefore, the second suction side check valve 30b blocks the reverse flow of the refrigerant gas that has passed through the compressor low-pressure sub-line 33 to the second compressor 14b. In this way, the low-pressure refrigerant gas can be recovered from the refrigerator 18 to the first compressor 14a through the low-pressure line 28. In addition, it is possible to prevent backflow from the second compressor 14 to the first compressor 14a through the low-pressure line 28. In addition, in the compressor 14, normally, the discharge port 20 and the suction port 21 are equalized when stopped. That is, both the discharge port 20 and the suction port 21 have an average pressure of high and low pressure (for example, if the high pressure is 2 MPa and the low pressure is 0.6 MPa, the average pressure is 1.3 MPa). Therefore, both the second discharge side check valve 29b and the second suction side check valve 30b have the outlet pressure significantly higher than the inlet pressure, and this pressure difference is surely closed. FIG. 3 is a control block diagram related to the cryopump system 10 of the embodiment. FIG. 3 shows the relevant part of the cryopump system 10, one of the plurality of cryopumps 12 shows internal details, and the other cryopumps 12 are omitted because they are the same. Similarly, the details of the plurality of compressors 14 are shown, and the other compressors 14 are the same, so the internal illustration is omitted. In addition, the control structure of this cryopump system 10 can be implemented on the hardware structure using components or circuits led by the computer's CPU or memory, and can be implemented on the software structure using computer programs, etc., in FIG. 3 Appropriately drawn as functional blocks realized through their cooperation. Those skilled in the art should understand that these functional blocks can be implemented in various forms by a combination of hardware and software. The cryopump system 10 includes a cryopump controller (hereinafter, also referred to as CP controller) 100. The CP controller 100 controls the cryopump 12 (ie, the refrigerator 18) and the compressor 14. The CP controller 100 includes a CPU that executes various arithmetic processes, a ROM that stores various control programs, a RAM used as a work area for storing data or executing programs, an input/output interface, and memory. The CP controller 100 is also configured to be able to communicate with a host controller (not shown) for controlling the vacuum program device in which the cryopump 12 is installed. The CP controller 100 is provided separately from the cryopump 12 and the compressor 14. The CP controller 100 is connected to the cryopump 12 and the compressor 14 so as to be able to communicate with each other. The cryopumps 12 each include an IO module 50 for input and output processing that communicates with the CP controller 100. In addition, the CP controller 100 may be integrally mounted on any cryopump 12 or compressor 14. As described above, the CP controller 100 is connected to the IO module 50 of each cryopump 12 in a communicable manner. The IO module 50 includes a refrigerator inverter 52 and a signal processing unit 54. The refrigerator 18 includes a refrigerator motor 56 as a driving source for driving the thermal cycle of the refrigerator 18. The refrigerator inverter 52 adjusts power of a predetermined voltage and frequency supplied from an external power source, such as a commercial power supply, and supplies it to the refrigerator motor 56. The CP controller 100 controls the voltage and frequency to be supplied to the refrigerator motor 56. The cryopump 12 includes a cryopanel temperature sensor 58. The cryopanel temperature sensor 58 measures the temperature of the cooling stage of the freezer 18 and/or the cryopanel 16 (see FIG. 1). The CP controller 100 determines the command control amount according to the sensor output signal. The signal processing unit 54 relays the command control amount sent from the CP controller 100 to the refrigerator inverter 52. For example, the signal processing unit 54 converts the command signal from the CP controller 100 into a signal that can be processed by the refrigerator inverter 52 and sends it to the refrigerator inverter 52. The command signal includes a signal indicating the operating frequency of the refrigerator 18. In addition, the signal processing unit 54 relays the outputs of various sensors of the cryopump 12 to the CP controller 100. For example, the signal processing unit 54 converts the sensor output signal into a signal that can be processed by the CP controller 100 and sends it to the CP controller 100. Various sensors including cryoplate temperature sensor 58 are connected to the signal processing section 54 of the IO module 50. The cryopanel temperature sensor 58 periodically measures the temperature of the cryopanel 16 and outputs a signal indicating the measured temperature value. The measured temperature signal of the cryopanel temperature sensor 58 is input to the CP controller 100 every predetermined time, and the temperature measurement value is stored and maintained in the CP controller 100 in a predetermined storage area. The operating frequency (also referred to as operating speed) of the freezer 18 represents the operating frequency or rotational speed of the freezer motor 56, the operating frequency of the freezer inverter 52, and the frequency of the thermal cycle of the freezer 18 (e.g., GM cycle, etc.) Or any of them. The frequency of the thermal cycle is the number of times per unit time of the thermal cycle performed in the refrigerator 18. The CP controller 100 is configured to determine the operating frequency of the refrigerator 18 as a function of the deviation between the target cooling temperature and the measured temperature of the cryopanel (for example, by PID control). The CP controller 100 outputs the determined operating frequency to the refrigerator inverter 52. The refrigerator inverter 52 converts the input power to have the operating frequency input from the CP controller 100. The input power to the refrigerator inverter 52 is supplied from a refrigerator power supply (not shown). The refrigerator inverter 52 outputs the converted electric power to the refrigerator motor 56. In this way, the refrigerator motor 56 is driven at the operating frequency determined by the CP controller 100 and output from the refrigerator inverter 52. When the thermal load on the cryopump 12 increases, the temperature of the cryopanel 16 may increase. In the case where the measured temperature of the cryopanel temperature sensor 58 is higher than the target temperature, the CP controller 100 increases the operating frequency of the refrigerator 18. As a result, the frequency of the thermal cycle in the freezer 18 also increases (that is, the freezing capacity of the freezer 18 increases), and the cryopanel 16 is cooled toward the target temperature. Conversely, when the measured temperature of the cryopanel temperature sensor 58 is lower than the target temperature, the operating frequency of the freezer 18 decreases, the freezing capacity decreases, and the cryopanel 16 heats up toward the target temperature. In this way, the temperature of the cryopanel 16 can be maintained within a temperature range around the target temperature. Since the operating frequency of the refrigerator 18 can be adjusted appropriately according to the heat load, such control helps to reduce the power consumption of the cryopump 12. In this way, the flow rate of the refrigerant gas used for cooling the cryopanel in the refrigerator 18 changes so as to maintain the temperature of the cryopanel 16 at the target temperature in response to the thermal load of the cryopump 12. Even if the flow rate of the refrigerant gas supplied from the compressor 14 to the refrigerator 18 is insufficient, even if the operating frequency of the refrigerator 18 increases, the freezing capacity of the refrigerator 18 will not increase sufficiently. Therefore, the cryopump system 10 of the embodiment is configured to make the refrigerant gas discharge flow rate from the compressor 14 variable. Hereinafter, an example of such compressor control will be described. In addition, the compressor 14 includes a compressor controller 60, a compressor inverter 62, a compressor motor 64, a first pressure sensor 66, and a second pressure sensor 68. The compressor 14 is configured as, for example, a scroll type, a rotary type, or another pump that boosts refrigerant gas, and a compressor motor 64 is provided as a driving source for driving the compressor 14. As in the case of the refrigerator 18, the compressor controller 60 determines the operating frequency of the compressor 14, and outputs the determined operating frequency to the compressor inverter 62. The compressor inverter 62 converts the input power according to the operation frequency input from the compressor controller 60 and outputs the converted power to the compressor motor 64. In this way, the compressor motor 64 is driven at the operating frequency determined by the compressor controller 60 and output from the compressor inverter 62. Here, the operating frequency of the compressor 14 refers to, for example, the operating frequency of the compressor inverter 62, the operating frequency or the rotation speed of the compressor motor 64. Inside the compressor 14, the first pressure sensor 66 is set to measure the high pressure of the cryopump system 10 (eg, the pressure of the high-pressure line 26), and the second pressure sensor 68 is set to measure the low pressure of the cryopump system 10 (For example, the pressure of the low-pressure line 28). The first pressure sensor 66 and the second pressure sensor 68 each periodically measure the pressure of the refrigerant gas, and output a signal indicating the measured pressure value to the compressor controller 60. The compressor controller 60 may send the measured pressure signal and/or the operating frequency of the compressor 14 to the CP controller 100. The compressor controller 60 is configured to control the operating frequency of the compressor 14 based on the measured pressure of the first pressure sensor 66 and/or the second pressure sensor 68. For example, the compressor controller 60 is configured to determine the operating frequency of the compressor 14 as a function of the deviation between the pressure difference between the discharge side and the suction side of the compressor 14 and the target pressure difference (for example, by PID control). Such control of the compressor 14 is sometimes referred to as "constant pressure difference control". In addition, if necessary, the target value of the differential pressure may be changed during the execution of the constant differential pressure control. In the constant pressure difference control, the compressor controller 60 obtains the pressure difference between the measured pressure of the first pressure sensor 66 and the measured pressure of the second pressure sensor 68. The compressor controller 60 determines the operating frequency of the compressor 14 such that the measured pressure difference matches the target pressure difference. If the measured pressure difference is greater than the target differential pressure value, the compressor controller 60 decreases the operating frequency, and if the measured pressure difference is less than the target differential pressure value, the operating frequency is increased. The flow rate of the refrigerant gas used in the cryogenic plate cooling by the refrigerator 18 is proportional to the operating frequency of the refrigerator 18, and is obtained, for example, by the product of the internal volume of the refrigerator 18 and the operating frequency of the refrigerator 18. As the operating frequency of the refrigerator 18 increases, the flow rate of the refrigerant gas to be supplied from the compressor 14 to the refrigerator 18 increases. At this time, if the operating frequency of the compressor 14 is low and the refrigerant gas supply from the compressor 14 is insufficient, the pressure on the discharge side of the compressor 14 decreases. As the operating frequency of the refrigerator 18 increases, the flow rate of refrigerant gas to be recovered from the refrigerator 18 to the compressor 14 also increases. At this time, if the operating frequency of the compressor 14 is low, the compressor 14 does not sufficiently recover the refrigerant gas discharged from the refrigerator 18, so the pressure on the suction side of the compressor 14 increases. As such, the increase in the operating frequency of the refrigerator 18 tends to reduce the pressure difference between the discharge side and the suction side of the compressor 14. Conversely, the decrease in the operating frequency of the refrigerator 18 causes the pressure difference between the discharge side and the suction side of the compressor 14 to tend to increase. According to the constant pressure difference control of the compressor 14, when the load on the cryopump 12 increases and the operating frequency of the refrigerator 18 increases, the operating frequency of the compressor 14 increases to suppress the pressure difference between the discharge side and the suction side of the compressor 14 When it decreases, the refrigerant gas supply from the compressor 14 to the refrigerator 18 also increases. On the other hand, when the load on the cryopump 12 decreases and the operating frequency of the refrigerator 18 decreases, the operating frequency of the compressor 14 decreases, and the supply of refrigerant gas from the compressor 14 to the refrigerator 18 is also suppressed. Since the operating frequency of the compressor 14 can be adjusted appropriately according to the load on the cryopump system 10, the constant pressure difference control helps reduce the power consumption of the cryopump system 10. However, a typical cryopump system can be configured with only one compressor. On the other hand, in the cryopump system 10 of the embodiment, not only one compressor 14 but also another compressor 14 is added. The cryopump system 10 has redundancy with respect to the compressor 14. These two compressors 14 are simultaneously operated as the refrigerant gas supply source to the cryopump 12 during the operation of the cryopump system 10. The total refrigerant gas supply capacity of the two compressors 14 is set to a total of the refrigerant gas flow rate required for each refrigerator 18 of the cryopump 12 to perform cryogenic plate cooling. Here, the refrigerant gas supply capacity of the compressor 14 refers to, for example, the maximum discharge flow rate of the compressor 14 achieved when the compressor 14 is operated at the maximum operating frequency. The refrigerant gas flow rate required by the refrigerator 18 refers to, for example, the refrigerant gas flow rate used in the refrigerator 18 when the refrigerator 18 is operated at the maximum operating frequency. Therefore, when the refrigerant gas supply capacities of the two compressors 14 are represented by Qc 1 and Qc 2 , and the refrigerant gas flow rates required by the two refrigerators 18 are represented by qr 1 and qr 2 , the following relationship holds. Qc 1 +Qc 2 ≥qr 1 +qr 2 By setting the refrigerant gas supply capacity of the compressor 14 in this way, the refrigerant gas can be sufficiently supplied to the two refrigerators 18 by the simultaneous operation of the two compressors 14. Since the shortage of refrigerant gas in the refrigerator 18 can be avoided, the cryopanel 16 can be maintained at the target temperature, and the operation of the cryopump system 10 can be continued. Furthermore, in the cryopump system 10 of the embodiment, regarding any compressor 14 of the two compressors 14, the refrigerant gas supply capacity of the compressor 14 is set to be not lower than that of the refrigerator 18 of each cryopump 12 to perform a cryogenic plate The total amount of refrigerant gas flow required for cooling. That is, the cryopump system 10 also satisfies the following relationship. Qc 1 ≥qr 1 +qr 2 and Qc 2 ≥qr 1 +qr 2 As explained with reference to FIG. 2, it is assumed that a compressor 14 is stopped for some reason. However, the cryopump system 10 can sufficiently supply refrigerant gas to the two refrigerators 18 by another compressor 14 that is not stopped. In this way, the cryopump system 10 can maintain the cryopanel 16 of each cryopump 12 at the target temperature even when one compressor 14 is not operating, and can continue the operation of the cryopump system 10. The structure of the cryopump system 10 of the embodiment can be generalized as follows. The cryopump system 10 includes M cryopumps 12 and N+1 compressors 14. The N+1 compressors 14 are connected in parallel and simultaneously operated to supply refrigerant gas to the refrigerators 18 of the cryopumps 12. Among them, M and N are positive integers. As an example, the positive integer M may be, for example, 1 or greater, 2 or greater, 3 or greater, 5 or greater, or 10 or greater. The positive integer M may be, for example, 20 or less, 10 or less, 5 or less, or 3 or less. The positive integer N may be, for example, 1 or greater, 2 or greater, 3 or greater, 5 or greater, or 10 or greater. The positive integer N may be, for example, 20 or less, 10 or less, 5 or less, or 3 or less. Regarding any N compressors 14 of the N+1 compressors 14, the total refrigerant gas supply capacity of the N compressors 14 is set to not less than the refrigerant required for the cryocooler 18 of each cryopump 12 to perform cryogenic plate cooling The total gas flow. Therefore, when the refrigerant gas supply capacity of the N+1 compressors 14 is represented by Qc 1 , Qc 2 ,..., Qc N , and Qc N+1 , the necessary refrigerant gas flow rate of the M refrigerators 18 is represented by qr 1 , When qr 2 ,..., qr M indicate, the cryopump system 10 satisfies all the following relationships. ΣQc-Qc 1 ≥Σqr ΣQc-Qc 2 ≥Σqr …… ΣQc-Qc N ≥Σqr ΣQc-Qc N+1 ≥Σqr, here, ΣQc=Qc 1 +Qc 2 +……+Qc N +Qc N+1 (That is, the total refrigerant gas supply capacity of N+1 compressors 14), Σqr=qr 1 +qr 2 +......+qr M (that is, the total refrigerant gas flow rate of the M refrigerators 18) . Therefore, the left side of the above equations shows the total refrigerant gas supply capacity of any N compressors 14 of the N+1 compressors 14. In this way, even if any compressor 14 is stopped for some reason, the cryopump system 10 can sufficiently supply refrigerant gas to the refrigerator 18 of the M cryopumps 12 by the remaining compressors 14 that are not stopped. . The cryopump system 10 can maintain the cryopanel 16 of each cryopump 12 at a target temperature while any compressor 14 is stopped, and can continue the operation of the cryopump system 10. In addition, under normal conditions in which all N+1 compressors 14 are operating, the cryopump system 10 includes one remaining compressor 14. Therefore, compared with the case where the cryopump system 10 includes only N compressors 14, the flow rate of refrigerant gas to be supplied to each of the N+1 compressors 14 may be smaller. Therefore, with the cryopump system 10 of the embodiment, each compressor 14 can be operated with a lower load (that is, operating frequency), which contributes to extending the life of the compressor 14. In addition, the cryopump system 10 includes a control unit that controls N+1 compressors 14 (for example, the compressor controller 60 or the CP controller 100). The control unit is configured to control each compressor 14 to increase the supply of refrigerant gas for each of the N compressors 14 operating simultaneously when the number of the compressors 14 operating simultaneously decreases from N+1 to N. An example suitable for such compressor control is the above-described constant pressure difference control. If one of the compressors 14 of the plurality of compressors 14 is stopped, because the total reduction of the refrigerant gas supply flow rate is equivalent to the amount of one of the compressors 14 being stopped, the pressure of the high-pressure line 26 may be reduced and the low pressure The pressure of line 28 increases. That is, if any of the compressors 14 stops, the pressure difference between the discharge side and the suction side of each of the remaining compressors 14 tends to decrease. According to the constant pressure difference control, the operating frequency of each compressor 14 is increased to restore the reduction of this kind of pressure difference to the target pressure difference. In this way, the cryopump system 10 can control each compressor 14 to increase the supply of refrigerant gas for each of the N compressors 14 operating simultaneously when the number of compressors 14 operating simultaneously decreases from N+1 to N . Furthermore, the piping system 24 of the cryopump system 10 includes a discharge-side check valve 29 and a suction-side check valve 30 for each compressor 14. In this way, even if any one of the plurality of compressors 14 is stopped, it is possible to prevent the backflow of refrigerant gas from the remaining compressor 14 in operation to the compressor 14 that is stopped. Since the discharge-side check valve 29 and the suction-side check valve 30 are mechanically closed by the pressure difference, the compressor 14 in stop can be naturally disconnected from the cryopump system 10 without electrical control. In particular, the discharge-side check valve 29 and the suction-side check valve 30 can use a general-purpose check valve that operates by the pressure difference between the inlet and the outlet. These check valves have a relatively simple structure and are inexpensive. The piping system 24 can be configured more simply than installing the electric control valve for disconnection in the piping system 24, which contributes to reducing the manufacturing cost of the cryopump system 10. In addition, if necessary, the piping system 24 may include a control valve instead of the discharge-side check valve 29 and/or the suction-side check valve 30, and the control valve is configured to block the reverse flow of the refrigerant gas to the compressor 14 that is stopped. In addition, the piping system 24 includes a set of detachable joints 35 provided on both sides of the discharge-side check valve 29. The piping system 24 includes another set of detachable joints 35 provided on both sides of the check valve 30 on the suction side. In this way, the operator can remove the stopped compressor 14 from the cryopump system 10 and perform maintenance. Alternatively, the operator can remove the compressor 14 from the cryopump system 10 and replace it with a new compressor or another compressor that has completed maintenance. Since it is possible to perform such maintenance work while continuing the operation of the cryopump system 10, it is convenient. The present invention has been described above based on the embodiments. The present invention is not limited to the above-mentioned embodiment, and various design changes can be made. Of course, those skilled in the art can understand that various modifications can be made, and these modifications are also included in the scope of the present invention. In addition, the various features described in association with one embodiment can also be applied to other embodiments. The new implementation forms generated by the combination also have the respective effects of the combined implementation forms. In the above embodiment, the refrigerant gas flow rate required by the refrigerator 18 refers to, for example, the refrigerant gas flow rate used in the refrigerator 18 when the refrigerator 18 is operated at the maximum operating frequency. Actually, the case where the maximum operating frequency of the freezer 18 is required is limited to when the cryopump system 10 is started (at this time, it is desired that the freezer 18 is cooled from room temperature at a high speed to an extremely low temperature) and the like is rare. As such, in a state where the cryopump system 10 is started and operates stably, the refrigerant gas flow rate required by the refrigerator 18 does not need to be large. Therefore, the refrigerant gas flow rate required by the refrigerator 18 may also refer to the refrigerant gas flow rate used in the refrigerator 18 when the refrigerator 18 is operated at a certain operating frequency threshold. The operating frequency threshold is less than the maximum operating frequency. In this way, the refrigerant gas supply capacity of the compressor 14 can be designed to be lower, and therefore the size of each compressor 14 can be reduced and the manufacturing cost of the cryopump system 10 can be reduced. The cryopump system 10 may also include at least one cryopump 12 and more than N+1 compressors 14 (eg, N+2 or N+3 compressors 14) operating simultaneously. With regard to any N compressors 14 out of the N+1 compressors 14, the total refrigerant gas supply capacity of the N compressors 14 is set to not less than each refrigerator 18 of at least one cryopump 12 The total flow of refrigerant gas required to perform cryoplate cooling. In this way, the cryopump system 10 is further redundant with respect to the compressor 14, for example, even if two or three compressors 14 have stopped, the operation of the cryopump system 10 can be continued. Alternatively, the remaining compressors 14 exceeding N+1 units may be installed in the cryopump system 10 as backup compressors that do not operate simultaneously with other compressors 14 under normal conditions.
